Control system for compression-ignition engine

ABSTRACT

A control system for a compression-ignition engine is provided, which includes an engine having a combustion chamber, an injector configured to supply fuel into the combustion chamber, a spark plug, a swirl valve provided to an intake passage of the engine, and a controller. The controller includes a processor configured to execute a swirl adjusting module to adjust a swirl valve opening to generate a swirl flow inside the combustion chamber, a fuel injection amount controlling module to control fuel injection amounts of pre-injection and post-injection so as to increase a ratio of an injection amount of the post-injection to a total fuel injection amount into the combustion chamber in one cycle as an engine speed increases, and a combustion controlling module to control the spark plug to ignite at a given ignition timing after the swirl generation and fuel injection, so that partial compression-ignition combustion is performed.

TECHNICAL FIELD

The present disclosure relates to a control system for acompression-ignition engine, which performs partial compression-ignitioncombustion in which a mixture gas within a cylinder is partiallycombusted by spark ignition (SI combustion) and then the remainingmixture gas within the cylinder is combusted by self-ignition (CI(compression ignition) combustion).

BACKGROUND OF THE DISCLOSURE

Recently, Homogeneous-Charge Compression Ignition (HCCI) combustion inwhich gasoline fuel mixed with air is combusted by self-ignition insidea sufficiently compressed combustion chamber has attracted attention.The HCCI combustion is a mode in which the mixture gas combusts at aplurality of positions simultaneously, and thus has a faster combustionspeed of the mixture gas than in SI combustion (spark ignitioncombustion) which is adopted for general gasoline engines. Therefore,the HCCI combustion is said to be significantly advantageous in terms ofthermal efficiency. However, the HCCI combustion has issues such as acombustion start timing of the mixture gas (a timing at which themixture gas self-ignites) greatly varying due to external factors (e.g.,atmospheric temperature) and control during a transient operation inwhich an engine load sharply changes being difficult.

Therefore, instead of combusting all of the mixture gas byself-ignition, it is proposed to combust a portion of the mixture gas byspark ignition using a spark plug. That is, after forcibly combusting aportion of the mixture gas through flame propagation caused by sparkignition (SI combustion), the remaining mixture gas is combusted byself-ignition (CI combustion). Hereinafter, such combustion is referredto as “partial compression-ignition combustion.”

For example, JP2009-108778A discloses an engine adopting a similarconcept to the partial compression-ignition combustion. This enginecauses flame propagation combustion by spark-igniting stratified mixturegas which is formed around a spark plug by a supplementary fuelinjection, and then performs a main fuel injection inside a combustionchamber warmed up by an effect of the flame propagation combustion, soas to combust the fuel injected in the main fuel injection throughself-ignition.

Although the engine of JP2009-108778A can stimulate CI combustion by thespark ignition using the spark plug, a state of a flame core formedimmediately after the spark ignition is considered to vary to someextent due to an environment of the combustion chamber. For example,when an engine speed is high, since a moving speed of a piston is high,the combustion chamber after the spark ignition rapidly expands to causeinsufficient growth of the flame core. When the growth of the flame coreis insufficient, a start timing of the CI combustion may greatly deviatefrom a target timing and the combustion may become unstable.

SUMMARY OF THE DISCLOSURE

The present disclosure is made in view of the above situations and aimsto provide a control system for a compression-ignition engine, whichsecures high combustion stability regardless of an engine speed.

In order to solve the above issue, according to one aspect of thepresent disclosure, a control system for a compression-ignition engineis provided. The device includes an engine having a combustion chamberformed by a cylinder, a piston and a cylinder head, an injector attachedto the engine and configured to supply fuel into the combustion chamber,a spark plug disposed to be oriented into the combustion chamber, aswirl valve provided to an intake passage of the engine, and acontroller connected to the injector, the spark plug and the swirlvalve, and configured to output a control signal to the injector, thespark plug and the swirl valve, respectively. The controller includes aprocessor configured to execute a swirl adjusting module to adjust anopening of the swirl valve to generate a swirl flow inside thecombustion chamber, a fuel injection amount controlling module tocontrol fuel injection amounts by the injector, of a pre-injection inwhich fuel is injected into the combustion chamber and a post-injectionin which fuel is injected again after the pre-injection, the fuelinjection amount controlling module outputting a control instruction tothe injector to increase a ratio of an injection amount of thepost-injection to a total amount of fuel to be injected into thecombustion chamber in one cycle as an engine speed increases, and acombustion controlling module to output an ignition instruction to thespark plug so as to ignite at a given ignition timing after the swirlgeneration by the swirl adjusting module and the fuel injection by thefuel injection amount controlling module, so that partialcompression-ignition combustion in which the mixture gas combusts byflame propagation and then combusts by compression ignition isperformed.

According to another aspect of the present disclosure, a control systemfor a compression-ignition engine is provided, the engine including aninjector configured to supply fuel into a combustion chamber and a sparkplug configured to ignite a mixture gas containing fuel supplied fromthe injector and air, the engine performing partial compression-ignitioncombustion including spark ignition (SI) combustion performed bycombusting a portion of the mixture gas through spark ignition by thespark plug, followed by compression ignition (CI) combustion performedby causing the remaining mixture gas to self-ignite. The device includesa swirl mechanism configured to generate a swirl flow, and a processorconfigured to, during the partial compression-ignition combustion,control the swirl mechanism to generate the swirl flow, control theinjector to perform a pre-injection in which a given amount of fuel isinjected into the combustion chamber and a post-injection in which fuelis injected again after the pre-injection, and increase a ratio of theinjection amount of the post-injection to a total amount of fuel to beinjected into the combustion chamber in one cycle as the engine speedincreases.

According to this configuration, during the partial compression-ignitioncombustion, since the ratio of the injection amount of thepost-injection to the total amount of fuel to be injected into thecombustion chamber in one cycle increases, a fuel concentration of themixture gas locally formed inside the combustion chamber at the ignitiontiming is increased when the engine speed is high. Thus, even under acondition where the engine speed is high and a rate of an expansioninside the combustion chamber after the spark ignition is high (as aresult, the flame core is difficult to grow), the formation and growthof the flame core are stimulated so that the SI combustion progressesstably and the subsequent CI combustion is surely performed, whichavoids the start timing of the CI combustion greatly varying betweencycles. As described above, according to this configuration, regardlessof the engine speed being high or low, the stable partialcompression-ignition combustion is achieved.

During the partial compression-ignition combustion in which the swirlflow is generated and the ratio of the injection amount of thepost-injection is increased as the engine speed increases, when theengine speed is higher than a given value, the fuel injection amountcontrolling module may control the injector to perform the pre-injectionand the post-injection on compression stroke.

According to this configuration, when the engine speed is high, comparedto a case where the pre-injection solely or both the pre-injection andthe post-injection is/are performed on the intake stroke, the mixturegas having the high fuel concentration is more reliably formed in thecombustion chamber. Thus, the flame core is formed and grown morereliably when the engine speed is high.

During the partial compression-ignition combustion in which the swirlflow is generated and the ratio of the injection amount of thepost-injection is increased as the engine speed increases, the fuelinjection amount controlling module may control the injector to performthe pre-injection on intake stroke and the pre-injection on compressionstroke when the engine speed is lower than a give value.

When the engine speed is low, expansion of the combustion chamber afterthe spark ignition is relatively slower, the flame core is easily formedand grown. Thus, even if the fuel amount supplied to the combustionchamber on the compression stroke is reduced by performing thepre-injection on the intake stroke, the partial compression-ignitioncombustion is suitably achieved when the engine speed is low.

Therefore, according to this configuration, when the engine speed islow, the mixture gas having the locally high fuel concentration isformed in the combustion chamber by the post-injection performed on thecompression stroke so that the partial compression-ignition combustionis achieved, and thermal efficiency improves by increasing the amount ofthe mixture gas for the CI combustion by diffusing the fuel wider by thepre-injection on the intake stroke.

The injector may at least have a first nozzle port and a second nozzleport disposed at a center portion of a ceiling surface of the combustionchamber and separated from each other in a circumferential direction ofthe injector. The first and second nozzle ports simultaneously injectthe fuel. The swirl flow may be an inclined swirl flow flowingnonparallel to a plane perpendicular to a center axis of the combustionchamber. The first and second nozzle ports may be positioned andoriented so that a first fuel portion injected by the first nozzle portthat has reached the swirl flow moves downstream along the swirl flowand then joins with a second fuel portion injected by the second nozzleport that has reached the swirl flow.

According to this configuration, by joining the fuels downstream of theswirl flow, the (rich) mixture gas at a relatively high fuelconcentration is formed in the center portion of the combustion chamberwhich is the final destination of the swirl flow. Thus, the flame coreis reliably formed and grows in the center portion of the combustionchamber.

During the partial compression-ignition combustion in which the swirlflow is generated and the ratio of the injection amount of thepost-injection is increased as the engine speed increases, the fuelinjection amount controlling module may control the injector so that theratio of an injection amount of the pre-injection to the total amount offuel to be injected into the combustion chamber in one cycle is reducedas the engine speed increases, and the ratio of the injection amount ofthe post-injection to the total amount of fuel to be injected into thecombustion chamber in one cycle increases as the engine speed increases.

In this manner, during the partial compression-ignition combustion, asthe engine speed increases, in addition to the ratio of the injectionamount of the post-injection, the ratio of the injection amount of alatter pre-injection performed at a retarding side in the pre-injectionis increased. Thus, when the engine speed is high, the fuelconcentration of the mixture gas locally formed in the combustionchamber at the ignition timing is increased furthermore reliably, andthe CI combustion is more surely performed.

During a partial compression-ignition combustion in an operating rangewith a higher load than that of an operating range where the partialcompression-ignition combustion in which the swirl flow is generated andthe ratio of the injection amount of the post-injection is increased asthe engine speed increases, the swirl adjusting module may cause theswirl valve to generate the swirl flow and the fuel injection amountcontrolling module controls the injector so that the pre-injection andthe post-injection are performed and the ratio of the injection amountof the post-injection to the total amount of fuel to be injected intothe combustion chamber in one cycle is reduced as the engine speedincreases.

According to this configuration, when the engine load and speed arehigh, the suitable CI combustion is achieved more surely. That is, whenthe engine load is high, the fuel amount injected into the combustionchamber increases, and when the engine speed is high, a time lengthcorresponding to one crank angle becomes short. Therefore, if the ratioof the fuel injection amount of the post-injection is increased when theengine load and speed are high, the fuel is not sufficiently vaporizeduntil the ignition timing, and soot, etc. may increase. Therefore,according to this configuration, when the engine load and speed arehigh, the suitable CI combustion is achieved while reducing thegeneration of soot, etc.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a system diagram schematically illustrating an overallconfiguration of a compression-ignition engine according to oneembodiment of the present disclosure.

FIG. 2 shows diagrams illustrating a cross-sectional view of ancombustion chamber and a plan view of a piston.

FIG. 3 is a schematic plan view illustrating a structure of a cylinderand intake and exhaust systems in the vicinity thereof.

FIG. 4 is a block diagram illustrating a control system of the engine.

FIG. 5 is a map in which operating ranges of the engine are dividedaccording to a difference in combustion mode.

FIG. 6 is a chart illustrating a waveform of a heat generation rate inSPCCI combustion (partial compression-ignition combustion).

FIG. 7 shows time charts schematically illustrating a combustion controlexecuted in each operating range of the engine.

FIG. 8 shows time charts illustrating an injection pattern within alow-load first segment.

FIG. 9 is a chart illustrating a relationship between an engine speedand an injection start timing within a first operating range.

FIG. 10 is a chart illustrating a relationship between the engine speedand an injection ratio timing within the first operating range.

FIG. 11 is a chart illustrating a relationship between the engine speedand an injection start timing within the first operating range.

FIG. 12 is a chart illustrating a relationship between the engine speedand an injection ratio timing within the first operating range.

FIG. 13 is a map illustrating a specific example of a target swirl valveopening set within the first operating range.

FIG. 14 is a chart illustrating a change in the target swirl valveopening when the engine speed is changed while the engine load is fixed.

FIG. 15 shows diagrams illustrating a behavior of fuel (spray of fuelportions) injected by an injector, in relation to a swirl flow.

FIG. 16 shows diagrams illustrating a mixture gas moving with the swirlflow within the combustion chamber seen from above.

FIG. 17 is a flowchart illustrating details of a control executed duringSPCCI combustion.

FIG. 18 is a chart corresponding to FIG. 6, illustrating various methodsof defining an SI ratio.

DETAILED DESCRIPTION OF THE DISCLOSURE (1) Overall Configuration ofEngine

FIGS. 1 and 2 are diagrams illustrating a suitable embodiment of acompression-ignition engine (hereinafter, simply referred to as “theengine”) to which a control system of the present disclosure is applied.The engine illustrated in FIGS. 1 and 2 is a four-cycle gasolinedirect-injection engine mounted on a vehicle as a drive source fortraveling, and includes an engine body 1, an intake passage 30 throughwhich intake air to be introduced into the engine body 1 flows, anexhaust passage 40 through which exhaust gas discharged from the enginebody 1 flows, and an external exhaust gas recirculation (EGR) device 50which recirculates a portion of the exhaust gas flowing through theexhaust passage 40 to the intake passage 30.

The engine body 1 has a cylinder block 3 formed therein with cylinders2, a cylinder head 4 attached to an upper surface of the cylinder block3 so as to cover above the cylinders 2, and a piston 5 reciprocatablyfitted into each cylinder 2. Typically, the engine body 1 is of amulti-cylinder type having a plurality of cylinders (e.g., fourcylinders). Here, the description is only given regarding one cylinder 2for the sake of simplicity.

A combustion chamber 6 is defined above the piston 5, and fuelcontaining gasoline as a main component is injected into the combustionchamber 6 by an injector 15 (described later). Further, the suppliedfuel is combusted while being mixed with air in the combustion chamber6, and an expansion force caused by this combustion pushes down thepiston 5 and thus it reciprocates in up-and-down directions of thecylinder. Note that the fuel injected into the combustion chamber 6 maybe any fuel as long as it contains gasoline as a main component and, forexample, it may contain a subcomponent, such as bioethanol, in additionto gasoline.

A crankshaft 7, which is an output shaft of the engine body 1, isprovided below the piston 5. The crankshaft 7 is connected to the piston5 via a connecting rod 8 and rotates about its center axis according tothe reciprocation (up-and-down motion) of the piston 5.

A geometric compression ratio of the cylinder 2, that is, a ratio of thevolume of the combustion chamber 6 when the piston 5 is at a top deadcenter (TDC) with the volume of the combustion chamber 6 when the piston5 is at a bottom dead center (BDC), is set between 13:1 and 30:1,preferably between 14:1 and 18:1, as a suitable value for SPCCIcombustion (partial compression-ignition combustion) described later.More specifically, the geometric compression ratio of the cylinder 2 isset between 14:1 and 17:1 in regular specifications using gasoline fuelhaving an octane number of about 91, and between 15:1 and 18:1 inhigh-octane specifications using gasoline fuel having an octane numberof about 96.

The cylinder block 3 is provided with a crank angle sensor SN1 whichdetects a rotational angle of the crankshaft 7 (crank angle) androtational speed (engine speed) of the crankshaft 7.

The cylinder head 4 is formed with an intake port 9 and an exhaust port10 which open into the combustion chamber 6, and provided with an intakevalve 11 which opens and closes the intake port 9, and an exhaust valve12 which opens and closes the exhaust port 10. Note that as illustratedin FIG. 2, the type of valve of the engine of this embodiment is afour-valve type including two intake valves and two exhaust valves. Thatis, the intake port 9 includes a first intake port 9A and a secondintake port 9B, and the exhaust port 10 includes a first exhaust port10A and a second exhaust port 10B. The intake valve 11 is provided toeach of the first and second intake ports 9A and 9B, and the exhaustvalve 12 is provided to each of the first and second exhaust ports 10Aand 10B.

As illustrated in FIG. 3, a swirl valve 18 configured to open and closethe second intake port 9B is provided therein. The swirl valve 18 isonly provided to the second intake port 9B and not to the first intakeport 9A. When such a swirl valve 18 is driven in the closing direction,since a rate of intake air flowing into the combustion chamber 6 fromthe first intake port 9A (in which the swirl valve 18 is not provided)increases, a circling flow circling around an axial line Z of thecylinder (a center axis of the combustion chamber 6), that is, the swirlflow, is enhanced. Conversely, driving the swirl valve 18 in the openingdirection weakens the swirl flow. Note that the intake port 9 of thisembodiment is a tumble port formable of a tumble flow (vertical vortex).Therefore, the swirl flow formed when closing the swirl valve 18 is aninclined swirl flow mixed with the tumble flow.

The intake valve 11 and the exhaust valve 12 are driven to open andclose in conjunction with the rotation of the crankshaft 7 by valveoperating mechanisms 13 and 14 including a pair of camshafts disposed inthe cylinder head 4.

The valve operating mechanism 13 for the intake valve 11 is builttherein with an intake variable valve timing mechanism (VVT) 13 aconfigured to change open and close timings of the intake valve 11.Similarly, the valve operating mechanism 14 for the exhaust valve 12 isbuilt therein with an exhaust VVT 14 a configured to change open andclose timings of the exhaust valve 12. The intake VVT 13 a (exhaust VVT14 a) is a so-called variable phase mechanism which simultaneouslychanges the open and close timings of the intake valve 11 (exhaust valve12) in the same amount. By controlling the intake VVT 13 a and theexhaust VVT 14 a, in this embodiment, a valve overlap period in whichboth the intake and exhaust valves 11 and 12 are opened over TDC ofexhaust stroke (ETDC) is adjusted, and by adjusting the valve overlapperiod, an amount of residual burned gas in the combustion chamber 6(internal EGR gas) is adjusted.

The cylinder head 4 is provided with the injector 15 which injects thefuel (mainly gasoline) into the combustion chamber 6, and a spark plug16 which ignites a mixture gas containing the fuel injected into thecombustion chamber 6 from the injector 15 and air introduced into thecombustion chamber 6. The cylinder head 4 is further provided with anin-cylinder pressure sensor SN2 which detects pressure of the combustionchamber 6 (hereinafter, also referred to as “in-cylinder pressure”).

As illustrated in FIG. 2, on a crown surface of the piston 5, a cavity20 is formed by denting a relatively wide area of the piston 5,including a center part thereof, to the opposite side from the cylinderhead 4 (downward). A center section of the cavity 20 is formed with abulge portion 20 a having a substantially conical shape, bulgingrelatively upward, and both sides of the cavity 20 over the bulgeportion 20 a in radial directions respectively form a bowl-shapedrecessed portion in cross section. In other words, the cavity 20 is arecessed portion having a donut shape in plan view, formed to surroundthe bulge portion 20 a. Further, a section of the crown surface of thepiston 5 radially outward of the cavity 20 is a squish portion 21comprised of an annular flat surface.

The injector 15 is disposed in a center portion of a ceiling surface ofthe combustion chamber 6, and its tip portion opposes to the centerportion (bulge portion 20 a) of the crown surface of the piston 5. Theinjector 15 is a multi-port injector having a plurality of nozzle portsat its tip portion. For example, the injector 15 has a total of tennozzle ports circumferentially arranged at an even interval, and theports simultaneously inject the fuel to spread radially (obliquelydownwardly). Lines L1 to L10 illustrated in FIG. 2 indicate center linesof the sprays of the fuel injected from the respective nozzle ports.When an angle at which each of the center lines L1 to L10 of the spraysintersects the cylinder axis Z is α, this intersecting angle α is30-60°, preferably approximately 45°. That is, when the injector 15injects the fuel, the sprays from the nozzle ports simultaneously spreadin directions at angles of 30-60° (preferably 45°) from the cylinderaxis Z.

With the total of ten nozzle ports provided in the injector 15 at theeven interval as in this embodiment, the center lines L1 to L10 of thesprays are arranged, 36° away from each other centering on the cylinderaxis Z. When a line extending perpendicular to the intake-exhaustdirection of the engine and extending perpendicular to the cylinder axisZ is a reference line K, the center lines L1 to L5 and the center linesL6 to L10 which are located opposite to each other with respect to thereference line K are in a line symmetric relationship to each other withrespect to the reference line K.

The spark plug 16 is located slightly off from the injector 15 to theintake side. The tip portion (electrode portion) of the spark plug 16 isset at a position overlapping with the cavity 20 in the plan view.

As illustrated in FIG. 1, the intake passage 30 is connected to one sidesurface of the cylinder head 4 to communicate with the intake port 9.Air (fresh air) taken in from an upstream end of the intake passage 30is introduced into the combustion chamber 6 through the intake passage30 and the intake port 9.

In the intake passage 30, an air cleaner 31 which removes foreignmatters within the intake air, a throttle valve 32 which adjusts a flowrate of intake air, a booster 33 which pumps the intake air whilecompressing it, an intercooler 35 which cools the intake air compressedby the booster 33, and a surge tank 36 are provided in order from theupstream side.

An airflow sensor SN3 which detects the flow rate of intake air, firstand second intake air temperature sensors SN4 and SN6 which detect atemperature of the intake air, and first and second intake air pressuresensors SN5 and SN7 which a detect pressure of the intake air areprovided in various parts of the intake passage 30. The airflow sensorSN3 and the first intake air temperature sensor SN4 are provided in aportion of the intake passage 30 between the air cleaner 31 and thethrottle valve 32, and detect the flow rate and the temperature of theintake air passing through this portion. The first intake air pressuresensor SN5 is provided in a portion of the intake passage 30 between thethrottle valve 32 and the booster 33 (downstream of a connection port ofan EGR passage 51 described later), and detects the pressure of theintake air passing through this portion. The second intake airtemperature sensor SN6 is provided in a portion of the intake passage 30between the booster 33 and the intercooler 35, and detects thetemperature of intake air passing through this portion. The secondintake air pressure sensor SN7 is provided in the surge tank 36 anddetects the pressure of intake air in the surge tank 36.

The booster 33 is a mechanical booster (supercharger) mechanicallylinked to the engine body 1. Although the specific type of the booster33 is not particularly limited, for example, any of known boosters, suchas Lysholm type, Roots type, or centrifugal type, may be used as thebooster 33.

An electromagnetic clutch 34 electrically switchable of its operationmode between “engaged” and “disengaged” is provided between the booster33 and the engine body 1. When the electromagnetic clutch 34 is engaged,a driving force is transmitted from the engine body 1 to the booster 33,and boosting by the booster 33 is performed. On the other hand, when theelectromagnetic clutch 34 is disengaged, the transmission of the drivingforce is interrupted, and the boosting by the booster 33 is stopped.

A bypass passage 38 which bypasses the booster 33 is provided in theintake passage 30. The bypass passage 38 connects the surge tank 36 tothe EGR passage 51 described later. A bypass valve 39 is provided in thebypass passage 38.

The exhaust passage 40 is connected to the other side surface of thecylinder head 4 so as to communicate with the exhaust port 10. Theburned gas generated in the combustion chamber 6 is discharged outsidethrough the exhaust port 10 and the exhaust passage 40.

A catalytic converter 41 is provided in the exhaust passage 40. Thecatalytic converter 41 is built therein with a three-way catalyst 41 awhich purifies hazardous components (HC, CO and NO_(x)) contained withinthe exhaust gas flowing through the exhaust passage 40, and a GPF(gasoline-particulate filter) 41 b which captures particulate matter(PM) contained within the exhaust gas. Note that another catalyticconverter built therein with a suitable catalyst, such as a three-waycatalyst or a NO_(x) catalyst, may be added downstream of the catalyticconverter 41.

The external EGR device 50 has the EGR passage 51 connecting the exhaustpassage 40 to the intake passage 30, and an EGR cooler 52 and an EGRvalve 53 which are provided in the EGR passage 51. The EGR passage 51connects a portion of the exhaust passage 40 downstream of the catalyticconverter 41 to a portion of the intake passage 30 between the throttlevalve 32 and the booster 33. The EGR cooler 52 cools the exhaust gasrecirculated from the exhaust passage 40 to the intake passage 30through the EGR passage 51 (external EGR gas) by heat exchange. The EGRvalve 53 is provided in the EGR passage 51 downstream of the EGR cooler52 (the side close to the intake passage 30), and adjusts the flow rateof the exhaust gas flowing through the EGR passage 51.

The EGR cooler 52 uses cooling water for cooling the engine body 1, as amedium (coolant) for heat exchange. The temperature of the external EGRgas to be recirculated to the intake passage 30 after being cooled bythe EGR cooler 52, although significantly drops compared to thetemperature of the exhaust gas immediately after being discharged fromthe combustion chamber 6, exceeds an outdoor temperature. Therefore,while the external EGR is executed, a compression start temperaturewhich is a temperature of the combustion chamber 6 when the compressionstroke is substantially started (the intake valve 11 is closed), becomeshigher than while the external EGR is not executed.

A pressure difference sensor SN8 which detects a difference betweenpressure upstream of the EGR valve 53 and pressure downstream thereof isprovided in the EGR passage 51.

(2) Control System

FIG. 4 is a block diagram illustrating a control system 60 of theengine. Some components of the engine are not illustrated in FIG. 4. AnECU (electronic control unit) 100 illustrated in FIG. 4 is amicroprocessor which comprehensively controls the engine, and comprisedof a well-known processor 101 such as a CPU having associated ROM, RAM,etc. The processor 101 is configured to execute a swirl adjusting module102 to control the swirl valve 18, a fuel injection amount controllingmodule 103 to control fuel injection amounts by the injector 15, and acombustion controlling module 104 to output an ignition instruction tothe spark plug 16 to ignite at a given ignition timing. These modulesare stored in non-transitory memory of the ECU 100 as software.

The ECU 100 receives detection signals from various sensors. Forexample, the ECU 100 is electrically connected to the crank angle sensorSN1, the in-cylinder pressure sensor SN2, the airflow sensor SN3, thefirst and second intake air temperature sensors SN4 and SN6, the firstand second intake air pressure sensors SN5 and SN7, and the pressuredifference sensor SN8, which are described above. The ECU 100sequentially receives the information detected by these sensors (i.e.,the crank angle, the engine speed, the in-cylinder pressure, the intakeair flow rate, the intake air temperatures, the intake air pressures,the difference in pressure between the upstream and downstream sides ofthe EGR valve 53, etc.).

Further, an accelerator sensor SN9 which detects an opening of anaccelerator pedal controlled by a vehicle driver driving the vehicle isprovided in the vehicle, and a detection signal from the acceleratorsensor SN9 is also inputted to the ECU 100.

The ECU 100 controls various components of the engine while executingvarious determinations and calculations based on the input signals fromthe various sensors. That is, the ECU 100 is electrically connected tothe intake VVT 13 a, the exhaust VVT 14 a, the injector 15, the sparkplug 16, the swirl valve 18, the throttle valve 32, the electromagneticclutch 34, the bypass valve 39, the EGR valve 53, etc., and outputscontrol signals to these components based on various calculationresults.

Note that the ECU 100 as described above corresponds to a “controller.”

(3) Control According to Operating State

FIG. 5 is a map illustrating a difference in control according to anengine speed and load. As illustrated in FIG. 5, an operating range ofthe engine is roughly divided into four operating ranges A1 to A4 due tothe difference in combustion mode. The fourth operating range A4 is ahigh-speed range in which the engine speed is high. The first operatingrange A1 is a low and medium-speed, low-load range in which the enginespeed is lower than the fourth operating range A4 and the engine load islow. The third operating range A3 is a low-speed high-load range inwhich the engine speed is low and the engine load is high. The secondoperating range A2 is a remaining range except for the first, third andfourth ranges A1, A3, and A4 (i.e., a range combining a low andmedium-speed, medium-load range and a medium-speed, high-load range).Hereinafter, the combustion mode, etc. selected in each operating rangewill be sequentially described.

(3-1) First Operating Range

Within the first operating range A1 in which the engine speed is low andthe engine load is low, the partial compression-ignition combustioncombining the SI combustion and the CI combustion (hereinafter referredto as “SPCCI combustion”) is performed. The SI combustion is a mode inwhich the mixture gas is ignited by the spark plug 16 and is thenforcibly combusted by flame propagation which spreads the combustingregion from the ignition point, and the CI combustion is a mode in whichthe mixture gas is combusted by self-ignition in an environmentincreased in temperature and pressure due to the compression of thepiston 5. The SPCCI combustion combining the SI combustion and the CIcombustion is a combustion mode in which the SI combustion is performedon a portion of the mixture gas inside the combustion chamber 6 by thespark ignition performed in an environment immediately before themixture gas self-ignites, and after the SI combustion, the CI combustionis performed on the remaining mixture gas in the combustion chamber 6 byself-ignition (by the further increase in temperature and pressureaccompanying the SI combustion). Note that “SPCCI” is an abbreviation of“SPark Controlled Compression Ignition.”

FIG. 6 is a chart illustrating a change in a heat generation rate(J/deg) according to the crank angle when the SPCCI combustion occurs.

The SPCCI combustion has a characteristic that the heat generation inthe CI combustion is faster than that in the SI combustion. For example,as illustrated in FIG. 6, a waveform of a heat generation rate caused bythe SPCCI combustion has a shape in which a rising slope in an earlystage of the combustion which corresponds to the SI combustion isgentler than a rising slope caused corresponding to the CI combustionoccurring subsequently. In other words, the waveform of the heatgeneration rate caused by the SPCCI combustion is formed to have a firstheat generation rate portion formed by the SI combustion and having arelatively gentle rising slope, and a second heat generation rateportion formed by the CI combustion and having a relatively sharp risingslope, which are next to each other in this order. Further,corresponding to the tendency of such a heat generation rate, in theSPCCI combustion, a pressure increase rate (dp/dθ) inside the combustionchamber 6 caused by the SI combustion is lower than that in the CIcombustion.

When the temperature and pressure inside the combustion chamber 6 risedue to the SI combustion, the unburned mixture gas self-ignites and theCI combustion starts. As illustrated in FIG. 6, the slope of thewaveform of the heat generation rate changes from gentle to sharp at thetiming of self-ignition (that is, the timing when the CI combustionstarts). That is, the waveform of the heat generation rate caused by theSPCCI combustion has a flection point at a timing when the CI combustionstarts (indicated by an “X” in FIG. 6).

After the CI combustion starts, the SI combustion and the CI combustionare performed in parallel. In the CI combustion, since the combustionspeed of the mixture gas is faster than that in the SI combustion, theheat generation rate becomes relatively high. However, since the CIcombustion is performed after TDC of compression stroke (CTDC), theslope of the waveform of the heat generation rate does not becomeexcessive. That is, after CTDC, since the motoring pressure decreasesdue to the piston 5 descending, the rise of the heat generation rate isprevented, which avoids excessive dp/dθ in the CI combustion. In theSPCCI combustion, due to the CI combustion being performed after the SIcombustion as described above, it is unlikely for dp/dθ which is anindex of combustion noise to become excessive, and combustion noise isreduced compared to performing the CI combustion alone (in the casewhere the CI combustion is performed on all the fuel).

The SPCCI combustion ends as the CI combustion finishes. Since thecombustion speed of the CI combustion is faster than that of the SIcombustion, the combustion end timing is advanced compared to performingthe SI combustion alone (in the case where the SI combustion isperformed on all the fuel). In other words, the SPCCI combustion bringsthe combustion end timing closer to CTDC, on the expansion stroke. Thus,the SPCCI combustion improves the fuel efficiency compared to the SIcombustion alone.

As specific modes of the SPCCI combustion, within the first operatingrange A1, a control for performing the SPCCI combustion of the mixturegas is performed while forming an environment in which an air-fuel ratio(A/F), which is a mass ratio of air (fresh air) to the fuel inside thecombustion chamber 6, is larger than a stoichiometric air-fuel ratio(14.7:1) (hereinafter, referred to as A/F lean environment). In order toachieve such the SPCCI combustion under the A/F lean environment, withinthe first operating range A1, the various components of the engine arecontrolled by the ECU 100 as follows.

The booster 33 is controlled to be OFF inside a boost line T illustratedin FIG. 5, and be ON outside the boost line T. Inside the boost line Twhere the booster 33 is OFF, i.e., at the lower speed side of the firstoperating range A1, the electromagnetic clutch 34 is disengaged todisconnect the booster 33 from the engine body 1 and the bypass valve 39is fully opened so as to stop boosting by the booster 33. Outside theboost line T where the booster 33 is ON, i.e., at the higher speed sideof the first operating range A1, the electromagnetic clutch 34 isengaged to connect the booster 33 to the engine body 1 so as to performboosting by the booster 33. Here, the opening of the bypass valve 39 iscontrolled so that the pressure in the surge tank 36 (boosting pressure)detected by the second intake air pressure sensor SN7 matches giventarget pressure determined for each operating condition (engine speedand engine load). For example, as the opening of the bypass valve 39increases, the flow rate of the intake air which flows back to theupstream side of the booster 33 through the bypass passage 38 increases,and as a result, the pressure of the intake air introduced into thesurge tank 36, that is, the boosting pressure, becomes low. By adjustingthe backflow amount of the intake air in this manner, the bypass valve39 controls the boosting pressure to the target pressure.

The intake VVT 13 a and the exhaust VVT 14 a set valve operation timingsof the intake and exhaust valves 11 and 12 so that the valve overlapperiod in which both the intake and exhaust valves 11 and 12 are openedover ETDC is formed. As a result, the internal EGR which leaves theburned gas inside the combustion chamber 6 is achieved, and thetemperature of the combustion chamber 6 is increased. That is, byopening the exhaust valve 12 until after ETDC (until the early stage ofthe intake stroke), the burned gas is drawn back into the combustionchamber 6 from the exhaust port 10 and the internal EGR is achieved. Thevalve overlap period (more specifically, the period in which the exhaustvalve 12 is opened on the intake stroke) is adjusted so that an internalEGR ratio which is a ratio of the internal EGR gas introduced into thecombustion chamber 6 increases as the engine load decreases.

Within the first operating range A1, a target internal EGR ratio whichis a target value of the internal EGR ratio is set to increasesubstantially between 10 and 50% as the engine load decreases. Note thatthe internal EGR ratio used here is a mass ratio of the residual burnedgas inside the combustion chamber 6 (internal EGR gas) to all the gasinside the combustion chamber 6. Further, the concept of the residualburned gas inside the combustion chamber 6 includes, not only the burnedgas residing inside the combustion chamber 6 without being discharged tothe exhaust port 10, but also the burned gas returned into thecombustion chamber 6 from the exhaust port 10 by opening the exhaustvalve 12 on the intake stroke.

The EGR valve 53 is opened in a major segment of the first operatingrange A1. Specifically, the EGR valve 53 is opened in the major segmentof the first operating range A1 which is other than its low speedsection, and the opening of the EGR valve 53 in this opened segment islarger as the engine speed is higher. Thus, an external EGR ratio whichis a ratio of exhaust gas recirculated to the combustion chamber 6through the EGR passage 51 (external EGR gas) is adjusted to increase asthe engine speed increases.

Within the first operating range A1, a target external EGR ratio whichis a target value of the external EGR ratio is set to increasesubstantially between 0 and 20% as the engine speed or load increases.Note that the external EGR ratio used here is a mass ratio of exhaustgas recirculated to the combustion chamber 6 through the EGR passage 51(external EGR gas) to all the gas inside the combustion chamber 6.

The throttle valve 32 is fully opened. Thus, a relatively large amountof air (fresh air) is introduced into the combustion chamber 6 and A/Fis set larger than the stoichiometric air-fuel ratio. In other words,within the first operating range A1, the SPCCI combustion is performedin the A/F lean environment in which an excess air ratio λ which is avalue obtained by dividing an actual air-fuel ratio by thestoichiometric air-fuel ratio is larger than 1. For example, the excessair ratio λ within the first operating range A1 is set to 2 or above sothat an amount of NO_(x) generated by the combustion is sufficientlyreduced. Note that in this embodiment, while the internal EGR and theexternal EGR are executed within the first operating range A1 asdescribed above, the amount of EGR gas introduced into the combustionchamber 6 by both the internal and external EGR (the internal EGR gasand the external EGR gas) needs to be set so that an amount of airequivalent to the target air-fuel ratio (λ>2) is secured inside thecombustion chamber 6. Respective target values of the internal EGR ratioand the external EGR ratio within the first operating range A1 aredetermined in advance to fulfill these requirements. That is, in thisembodiment, the target values of the internal EGR ratio and the externalEGR ratio are set so that an amount of gas obtained by subtracting anair amount equivalent to the target air-fuel ratio (λ>2) from the totalgas amount introduced into the combustion chamber 6 in the state wherethe throttle valve 32 is fully opened is introduced into the combustionchamber 6 as the internal EGR gas and the external EGR gas. Further, thevalve overlap period and the opening of the EGR valve 53 arerespectively adjusted according to the target values of the EGR ratios.

The opening of the swirl valve 18 is set smaller than a half-openedstate (50%). By reducing the opening of the swirl valve 18, a majorportion of the intake air introduced into the combustion chamber 6 issucked in from the first intake port 9A (the intake port to which theswirl valve 18 is not provided), and a strong swirl flow is formedinside the combustion chamber 6. This swirl flow grows on the intakestroke, remains until an intermediate section of the compression stroke,and stimulates stratification of the fuel. That is, a concentrationdifference that the fuel concentration is higher in the center portionof the combustion chamber 6 than outside thereof (outer circumferentialportion) is formed. Although described in detail in Section (4-2) later,within the first operating range A1, the air-fuel ratio in the centerportion of the combustion chamber 6 is set to between 20:1 and 30:1 andthe air-fuel ratio in an outer circumferential portion of the combustionchamber 6 is set to 35:1 or above. Further, the opening of the swirlvalve 18 is reduced to be smaller as the engine speed is lower. Thus,the intensity of the swirl flow is adjusted to be higher as the enginespeed is lower.

The injector 15 injects the fuel by splitting it into a plurality ofinjections from the intake stroke to the compression stroke. Further thespark plug 16 ignites the mixture gas near CTDC. For example, within thefirst operating range Al, the spark plug 16 ignites the mixture gas at aslightly advanced timing from CTDC. This ignition triggers the SPCCIcombustion, a portion of the mixture gas in the combustion chamber 6 iscombusted through flame propagation (SI combustion), and then theremaining mixture gas is combusted by self-ignition (CI combustion). Thedetails of the injection pattern in the first operating range A1 will bedescribed later.

(3-2) Second Operating Range

FIG. 7 is a chart illustrating an injection pulse of the injector 15,the ignition timing (the timing when the spark plug 16 ignites themixture gas) and the heat generation rate at an operation point P4within the second operating range A2, an operation point P5 within thethird operating range A3, and an operation point P6 within the fourthoperating range A4. The width of the injection pulse indicates theinjection amount, and the injection amount is larger as the width of theinjection pulse increases.

Within the second operating range A2 (a range combined the low andmedium-speed, medium-load range and the medium-speed, high-load range),a control for performing the SPCCI combustion of the mixture gas isperformed while forming an environment in which a gas air-fuel ratio(G/F), which is a ratio of all the gas to the fuel inside the combustionchamber 6, is larger than the stoichiometric air-fuel ratio (14.7:1) andA/F substantially matches the stoichiometric air-fuel ratio(hereinafter, referred to as G/F lean environment), is executed. Forexample, in order to achieve the SPCCI combustion in such a G/F leanenvironment, within the second operating range A2, various components ofthe engine are controlled by the ECU 100 as follows.

The injector 15 injects at least a portion of the fuel to be injected inone combustion cycle, during the compression stroke. For example, at anoperation point P4 within the second operating range A2, the injector 15injects the fuel separately in two times in an early half and latterhalf of the compression stroke, as illustrated in Part (a) of FIG. 7.

The spark plug 16 ignites the mixture gas near CTDC. For example, at theoperation point P4, the spark plug 16 ignites the mixture gas at aslightly advanced timing from CTDC. This ignition triggers the SPCCIcombustion, and a portion of the mixture gas inside the combustionchamber 6 is combusted through flame propagation caused by sparkignition (SI combustion), then the remaining mixture gas is combusted byself-ignition (CI combustion).

The booster 33 is controlled to be OFF in a section of the low-load andlow-speed range overlapping with the section inside the boost line T,and be ON outside this section. When the booster 33 is ON and boostingthe intake air, the opening of the bypass valve 39 is controlled so thatthe pressure inside the surge tank 36 (boosting pressure) matches withthe target pressure.

The intake VVT 13 a and the exhaust VVT 14 a set the valve operationtimings of the intake and exhaust valves 11 and 12 so that the valveoverlap period of a given length is formed. Note that since the boostingis performed (i.e., the intake air pressure is increased) withinsubstantially the entire second operating range A2, even when theexhaust valve 12 is opened on the intake stroke, the backflow of theburned gas into the combustion chamber 6 from the exhaust port 10 (i.e.,internal EGR) does not easily occur. Thus, the internal EGR ratio withinthe second operating range A2 is smaller than that within the firstoperating range A1, and the internal EGR is substantially stoppedparticularly in the higher load side of the second operating range A2.

The throttle valve 32 is fully opened.

The opening of the EGR valve 53 is controlled so that the air-fuel ratio(A/F) in the combustion chamber 6 becomes the stoichiometric air-fuelratio (λ=1) or therearound. For example, the EGR valve 53 adjusts theamount of the exhaust gas recirculated through the EGR passage 51(external EGR gas) so that the excess air ratio λ becomes 1±0.2. Notethat since the air amount equivalent to the stoichiometric air-fuelratio increases as the engine load increases, accordingly the externalEGR ratio within the second operating range A2 is set to be smaller asthe engine load increases (in other words, it is set to be larger as theengine load decreases). The opening of the EGR valve 53 is controlledaccording to the target value of the external EGR ratio set in thismanner.

The opening of the swirl valve 18 is set substantially the same aswithin the first operating range A1 or a given intermediate openinglarger than this opening.

(3-3) Third Operating Range

Within the third operating range A3 on the low-speed and high-load side,a control is executed in which at least a portion of the fuel isinjected in the final stage of the compression stroke and the mixturegas is subjected to the SI combustion. For example, in order to achievethe SI combustion accompanied by such a retarded injection, within thethird operating range A3, the various components of the engine arecontrolled by the ECU 100 as follows.

The injector 15 injects at least a portion of the fuel to be injected inone combustion cycle in the final stage of the compression stroke. Forexample, at an operation point P5 included in the third operating rangeA3, as illustrated in Part (b) of FIG. 7, the injector 15 injects allthe fuel to be injected in one cycle in the final stage of thecompression stroke (immediately before CTDC).

The spark plug 16 ignites the mixture gas at a relatively retardedtiming, for example 5° CA to 20° CA from CTDC. Further, this ignitiontriggers the SI combustion, and all the mixture gas in the combustionchamber 6 combusts through flame propagation. Note that the reason whythe ignition timing within the third operating range A3 is retarded asdescribed above is to prevent abnormal combustion, such as knocking andpre-ignition. However, within the third operating range A3, the fuelinjection is set to be performed in the final stage of the compressionstroke (immediately before CTDC), which is considerably late, therefore,even with the ignition timing retarded as described above, thecombustion speed after the ignition (flame propagation speed) isrelatively fast. That is, since the period from the fuel injection tothe ignition is sufficiently short, the flow (turbulence kinetic energy)in the combustion chamber 6 at the ignition timing becomes relativelystrong, and the combustion speed after the ignition is accelerated usingthis flow. Thus, thermal efficiency is kept high while preventingabnormal combustion.

The booster 33 is controlled to be ON and performs boosting. Theboosting pressure here is adjusted by the bypass valve 39.

The throttle valve 32 is fully opened.

The opening of the EGR valve 53 is controlled so that the air-fuel ratio(A/F) in the combustion chamber 6 becomes the stoichiometric air-fuelratio (λ=1) or therearound. For example, the EGR valve 53 adjusts theamount of the exhaust gas recirculated through the EGR passage 51(external EGR gas) so that the excess air ratio λ becomes 1±0.2.

The opening of the swirl valve 18 is set to or near a half-opened state(50%).

(3-4) Fourth Operating Range

Within the fourth opening range A4 on the higher speed side of the firstto third operating ranges A1 to A3, relatively basic SI combustion isexecuted. In order to achieve this SI combustion, within the fourthoperating range A4, the various components of the engine are controlledby the ECU 100 as follows.

The injector 15 at least injects the fuel over a given periodoverlapping with the intake stroke. For example, at an operation pointP6 within the fourth operating range A4, the injector 15 injects thefuel over a continuous period from the intake stroke to the compressionstroke, as illustrated in Part (c) of FIG. 7. Note that since theoperation point P6 corresponds to a considerably high-speed andhigh-load condition, the amount of fuel to be injected in one combustioncycle is large and also a crank angle period required for injecting therequired amount of fuel becomes long, for which the fuel injectionperiod at the operation point P6 is longer than the other operationpoints (P4, P5) described above.

The spark plug 16 ignites the mixture gas near CTDC. For example, at theoperation point P6, the spark plug 16 ignites the mixture gas at aslightly advanced timing from CTDC. Further, this ignition triggers theSI combustion, and all the mixture gas in the combustion chamber 6combusts through flame propagation.

The booster 33 is controlled to be ON and performs boosting. Theboosting pressure here is adjusted by the bypass valve 39.

The throttle valve 32 is fully opened.

The opening of the EGR valve 53 is controlled so that the air-fuel ratio(A/F) in the combustion chamber 6 becomes the stoichiometric air-fuelratio or slightly richer (λ≤1).

The swirl valve 18 is fully opened. Thus, not only the first intake port9A but also the second intake port 9B are fully opened and chargingefficiency of the engine is improved.

(4) Injection Pattern within First Operating Range

Next, the details of the fuel injection control within the firstoperating range A1 will be described. The injection pattern with respectto the engine speed changes differently between the low load side andthe high load side within the first operating range A1. Hereinafter, alow engine load segment of the first operating range A1 is referred toas a low-load first segment A1_a, and a high engine load segment thanthe low-load first segment A1_a within the first operating range A1 isreferred to as a high-load first segment A1_b. For example, asillustrated in FIG. 5, the low-load first segment A1_a is a segmentwhere the engine load is lower than a given reference load T2 within thefirst operating range A1, and the high-load first segment A1_b is asegment where the engine load is higher than the given reference load T2within the first operating range A1.

FIG. 8 is a chart illustrating the injection pulse of the injector 15,the ignition timing and the heat generation rate at the operation pointswithin the first operating range A1. The width of the injection pulse(i.e., injection period) indicates the substantial injection amount andthe injection amount is larger as the width of the injection pulseincreases.

Part (a) of FIG. 8 illustrates an injection pattern of the fuel, etc.within an operation point P1_a where the engine speed is low in thelow-load first segment A1_a and an operation point P1_b where the enginespeed is low in the high-load first segment A1_b. That is, at anoperation point where the engine speed is low within the first operatingrange A1, the fuel injection pattern and the ignition timing are fixedregardless of the engine load.

As illustrated in Part (a) of FIG. 8, at the operation points P1_a andP1_b, a total of three injections are performed in one cycle. Apre-injection Q1 which is performed first and a middle injection Q2performed second are carried out from an early stage to an intermediatestage of the intake stroke. A post injection Q3 performed thirdly iscarried out in a final stage of the compression stroke. In thisspecification, the early, intermediate, and final stages of a certainstroke mean respective stages when the certain stroke is divided intothree stages.

As illustrated in Part (a) of FIG. 8, at the operation points P1_a andP1_b, a major portion of the fuel for one cycle is injected in thepre-injection Q1 and the remaining small amount of fuel is injected inthe intermediate and post injections Q2 and Q3. That is, at theoperation points P1_a and P1_b, a ratio of the injection amount of thepre-injection Q1 to the total amount of fuel to be injected in one cycle(hereinafter, simply referred to as “the ratio of the injection amountof the pre-injection Q1”) is larger than a ratio of the injection amountof the middle injection Q2 to the total amount of fuel to be injected inone cycle (hereinafter, simply referred to as “the ratio of theinjection amount of the middle injection Q2”), and a ratio of theinjection amount of the post injection Q3 to the total amount of fuel tobe injected in one cycle (hereinafter, simply referred to as “the ratioof the injection amount of the post injection Q3”). For example, theratio of the injection amount of the pre-injection Q1 is about 80%, andthe ratio of the injection amount of the middle injection Q2 and theratio of the injection amount of the post injection Q3 are respectivelyabout 10%.

The fuel injection amount in the pre-injection Q1 is at least 50% of thefuel to be injected into the combustion chamber 6 in one cycle.

Thus, at the operation points P1_a and P1_b, the timing when the fuel isinjected into the combustion chamber 6, more specifically, a completiontiming of the injection of 50% amount of the fuel into the combustionchamber 6 in one cycle, which is an average timing of the fuel injectioninto the combustion chamber 6 (hereinafter, suitably referred to as “thecenter timing of the fuel injection”), is a timing within the injectionperiod of the pre-injection Q1 (a period during which the fuel isinjected). That is, at the operation points P1_a and P1_b, the centertiming of the fuel injection is a given timing within the injectionperiod of the pre-injection Q1, on the intake stroke.

Part (b) of FIG. 8 illustrates an injection pattern of the fuel, etc. atan operation point P2 where the engine speed is high in the low-loadfirst segment A1_a.

As illustrated in Part (b) of FIG. 8, at the operation point P2, similarto the operation point P1_a, a total of three injections are performedin one cycle. Note that at the operation point P2, different from theoperation point P1_a, all the pre-injection Q1 performed firstly, themiddle injection Q2 performed secondly, and the post injection Q3performed thirdly are carried out from the intermediate stage to thefinal stage of the compression stroke. That is, at the operation pointP2, all the fuel for one cycle is injected on the compression stroke.Thus, at the operation point P2, the center timing of the fuel injectionis on the compression stroke.

As described above, at the operation point P1_a, the center timing ofthe fuel injection is on the intake stroke. Therefore, the center timingof the fuel injection at the operation point P2 is retarded than that atthe operation point P1_a.

As illustrated in Part (b) of FIG. 8, at the operation point P2, thepost injection Q3 contains a largest injection amount, and the injectionamount of the pre-injection Q1 and the injection amount of the middleinjection Q2 are substantially the same as each other. For example, theratio of the injection amount of the post injection Q3 is about 45%, theratio of the injection amount of the pre-injection Q1 is about 25%, andthe ratio of the injection amount of the middle injection Q2 is about30%.

As described above, at the operation point P1_a, a major portion of thefuel for one cycle is injected in the pre-injection Q1, and the ratio ofthe injection amount of the post injection Q3 is extremely small (e.g.,about 10%). In this regard, at the operation point P2, the postinjection Q3 contains a largest injection amount, and the ratio of theinjection amount of the post injection Q3 is larger than that at theoperation point P1_a.

Part (c) of FIG. 8 illustrates an injection pattern of the fuel, etc.within an operation point P3 where the engine speed is high in thehigh-load first segment A1_b.

As illustrated in Part (c) of FIG. 8, at the operation point P3, evenwithin the same high-load first segment A1_b, different from theoperation point P1_b where the engine speed is low, the post injectionQ3 is not performed and a total of two injections are performed in onecycle. That is, at the operation point P3, only the pre-injection Q1 andthe middle injection Q2 are performed.

Also at the operation point P3, similar to the operation point P1_b, thepre-injection Q1 performed firstly is carried out from the early stageto the intermediate stage of the intake stroke. On the other hand, atthe operation point P3, different from the operation P1_b, the middleinjection Q2 performed secondly is carried out from the intermediatestage to the final stage of the compression stroke.

As illustrated in Part (c) of FIG. 8, at the operation point P3, a majorportion of the fuel for one cycle is injected in the pre-injection Q1and the remaining small amount of fuel is injected in the middleinjection Q2. That is, at the operation point P3, the ratio of theinjection amount of the pre-injection Q1 is set larger than the ratiosof the injection amounts of the intermediate and post injections Q2 andQ3. In this embodiment, as described above, at the operation point P3,the post injection Q3 is not performed and the ratio of the injectionamount of the post injection Q3 is 0%. For example, the ratios of theinjection amounts of the pre-injection Q1 and the middle injection Q2are about 90% and 10%, respectively.

The fuel injection amount in the pre-injection Q1 is at least 50% of thefuel to be injected into the combustion chamber 6 in one cycle.Accordingly, at the operation point P3, the center timing of the fuelinjection is in the injection period of the pre-injection Q1, i.e., atiming on the intake stroke.

(Relationship Between Engine Speed and Injection Pattern Within Low-LoadFirst Segment A1_a) (Injection Start Timing)

FIG. 9 is a chart illustrating a change in the injection start timing ofeach of the injections Q1, Q2 and Q3 in relation to the engine speed, inthe low-speed first segment A1_a. FIG. 9 also illustrates the operationpoints P1_a and P2 described above.

As illustrated in FIG. 9, in the low-load first segment A1_a as a whole,the injection start timing of the pre-injection Q1 is set to retard asthe engine speed increases. Further, in the low-load first segment A1_a,the pre-injection Q1 is performed on the intake stroke and the postinjection Q3 is performed on the compression stroke on the lower enginespeed side, whereas the all injections Q1, Q2 and Q3 are performed onthe compression stroke on the higher engine speed side.

Specifically, within an engine speed range below a first speed N1, theinjection start timing of the pre-injection Q1 is set to a timing ti_1on the intake stroke, fixed regardless of the engine speed. When theengine speed exceeds the first speed N1, the injection start timing ofthe pre-injection Q1 is gradually regarded from the timing ti_1 for whenthe engine speed is the first speed N1, as the engine speed increases.At an engine speed N10 between the first speed N1 and a second speed N2,the injection start timing of the pre-injection Q1 is the bottom deadcenter of the intake stroke (IBDC). Even when the engine speed exceedsthe engine speed N10, the injection start timing of the pre-injection Q1is gradually regarded as the engine speed increases. When the enginespeed becomes the second speed N2, the injection start timing of thepre-injection Q1 is set to a timing ti_2 which is retarded than IBDC andis on the compression stroke. Within an engine speed range above thesecond speed N2, the injection start timing of the pre-injection Q1 isset to the timing ti_2 for when the engine speed is the second speed N2,fixed regardless the engine speed.

As illustrated in FIG. 9, in the low-load first segment A1_a, similar tothe injection start timing of the pre-injection Q1, the injection starttiming of the middle injection Q2 is set to retard as the engine speedincreases.

Specifically, within the engine speed range below the first speed N1,the injection start timing of the middle injection Q2 is set to a timingti_3 on the intake stroke, fixed regardless of the engine speed. Whenthe engine speed exceeds the first speed N1, the injection start timingof the middle injection Q2 is gradually retarded from the timing ti_3 asthe engine speed increases. Within the engine speed range above thesecond speed N2, the injection start timing of the middle injection Q2is set to a timing ti_4 for when the engine speed is the second speedN2, fixed regardless of the engine speed (a timing retarded from thetiming ti_3 for when the engine speed is the first speed N1). Theinjection start timing of the middle injection Q2 is also set to IBDC atan engine speed between the first speed N1 and the second speed N2, andwhen the engine speed is the second speed N2, the injection start timingof the middle injection Q2 is on the compression stroke.

On the other hand, as illustrated in FIG. 9, in the low-load firstsegment A1_a, the injection start timing of the post injection Q3 is setto advance as the engine speed increases.

Specifically, within the engine speed range below the first speed N1,the injection start timing of the post injection Q3 is set to a timingti_5 on the compression stroke, fixed regardless of the engine speed.When the engine speed exceeds the first speed N1, the injection starttiming of the post injection Q3 gradually advances from the timing ti_5as the engine speed increases. Within the engine speed range above thesecond speed N2, the injection start timing of the post injection Q3 isset to the timing ti_6 for when the engine speed is the second speed N2,fixed regardless of the engine speed (a timing advanced from the timingti_5 for when the engine speed is the first speed N1).

(Ratio of Injection Amount)

FIG. 10 is a chart illustrating a change in the ratio of the injectionamount of each of the injections Q1, Q2, and Q3 in relation to theengine speed, in the low-load first segment A1_a.

As illustrated in FIG. 10, in the low-load first segment A1_a as awhole, the ratio of the injection amount of the pre-injection Q1 is setsmaller as the engine speed increases.

Specifically, in the engine speed range below the first speed N1, theratio of the injection amount of the pre-injection Q1 is fixedregardless of the engine speed and is set to a ratio Ri_1 close to 100%.When the engine speed exceeds the first speed N1, the ratio of theinjection amount of the pre-injection Q1 gradually decreases from theratio Ri_1 as the engine speed increases. Within the engine speed rangeabove the second speed N2, the ratio of the injection amount of thepre-injection Q1 is a ratio Ri_2 for when the engine speed is the secondspeed N2, fixed regardless of the engine speed (a value smaller than theratio Ri_1 for when the engine speed is the first speed N1). Forexample, as described above, the ratio Ri_1 of the injection amount ofthe pre-injection Q1 at the operation point P1_a when the engine speedis below the first speed N1 is about 80%, and the ratio Ri_2 of theinjection amount of the pre-injection Q1 at the operation point P2 whenthe engine speed is above the second speed N2 is about 30%.

On the other hand, as illustrated in FIG. 10, in the low-load firstsegment A1_a as a whole, the ratio of the injection amount of the middleinjection Q2 is set to be larger as the engine speed increases.

Specifically, in the engine speed range below the first speed N1, theratio of the injection amount of the middle injection Q2 is fixedregardless of the engine speed and is set to a value Ri_3 close to 0.When the engine speed exceeds the first speed N1, the ratio of theinjection amount of the middle injection Q2 gradually increases from theratio Ri_3 as the engine speed increases. Within the engine speed rangeabove the second speed N2, the ratio of the injection amount of themiddle injection Q2 is a ratio Ri_4 for when the engine speed is thesecond speed N2, fixed regardless of the engine speed (a value largerthan the ratio Ri_3 for when the engine speed is the first speed N1).For example, as described above, the ratio Ri_3 of the injection amountof the middle injection Q2 at the operation point P1_a when the enginespeed is the first speed N1 is about 10%, and the ratio Ri_4 of theinjection amount of the middle injection Q2 at the operation point P2when the engine speed is higher than the second speed N2 is about 25%.

As illustrated in FIG. 10, in the low-load first segment A1_a as awhole, the ratio of the injection amount of the post injection Q3 is setlarger as the engine speed increases.

Specifically, in the engine speed range below the first speed N1, theratio of the injection amount of the post injection Q3 is fixedregardless of the engine speed and is set to a value Ri_5 close to 0.When the engine speed exceeds the first speed N1, the ratio of theinjection amount of the post injection Q3 gradually increases from theratio Ri_5 as the engine speed increases. Within the engine speed rangeabove the second speed N2, the ratio of the injection amount of the postinjection Q3 is a ratio Ri_6 for when the engine speed is the secondspeed N2, fixed regardless of the engine speed (a value larger than theratio Ri_5 for when the engine speed is the first speed N1). Forexample, as described above, the ratio Ri_5 of the injection amount ofthe post injection Q3 at the operation point P1_a when the engine speedis the first speed N1 is about 10%, and the ratio Ri_6 of the injectionamount of the post injection Q3 at the operation point P2 when theengine speed is higher than the second speed N2 is about 45%.

Further as illustrated in FIG. 10, in the low-load first segment A1_a,when the engine speed is lower than an intermediate speed N11 which isbetween the first speed N1 and the second speed N2, the ratio of theinjection amount of the pre-injection Q1 has the largest value, and whenthe engine speed is higher than the intermediate speed N11, the ratio ofthe injection amount of the post injection Q2 has the largest value,among the values Q1, Q2, and Q3.

By controlling the injection start timings and the injection amounts ofthe injections Q1, Q2, and Q3 as described above, in the low-load firstsegment A1_a, the center timing of the fuel injection is retarded as theengine speed increases.

(Relationship Between Engine Speed and Injection Pattern WithinHigh-Load First Segment A1_b) (Injection Start Timing)

FIG. 11 is a chart illustrating a change in the injection start timingof each of the injections Q1, Q2, and Q3 in relation to the enginespeed, in the high-load first segment A1_b.

As illustrated in FIG. 11, in the high-load first segment A1_b, as awhole, the injection start timing of the pre-injection Q1 issubstantially fixed regardless of the engine speed.

Specifically, within an engine speed range below the first speed N1, theinjection start timing of the pre-injection Q1 is set to a timing ti_11on the intake stroke, fixed regardless of the engine speed. When theengine speed exceeds the first speed N1, the injection start timing ofthe pre-injection Q1 is gradually retarded as the engine speedincreases, and in the engine speed range above the second speed N2, itis gradually advanced as the engine speed increases. When exceeding thethird speed N3, which is greater than the second speed N2, the injectionstart timing of the pre-injection Q1 is set to a timing ti_13 fixedregardless of the engine speed. As described above, in the high-loadfirst segment A1_b, the injection start timing of the pre-injection Q1is set to be the most-retarded timing ti_12 (of the timings ti_11,ti_12, and ti_13 of the pre-injection Q1) when the engine speed is thesecond speed N2. Note that a difference between the injection starttiming ti_11 of the pre-injection Q1 when the engine speed is the firstspeed N1 and the injection start timing ti_12 of the pre-injection Q1 atthe second speed N2 is extremely small (e.g., this difference is 20° CAor below), a difference between the injection start timing ti_13 of thepre-injection Q1 when the engine speed is the third speed N3 and theinjection start timing ti_12 of the pre-injection Q1 at the second speedN2 is extremely small (e.g., this difference is 15° CA or below), and inthe high-load first segment A1_b, the injection start timing of thepre-injection Q1 is substantially fixed regardless of the engine speed.At least a change width of the injection start timing of thepre-injection Q1 in the high-load first segment A1_b with respect to theengine speed is sufficiently smaller than a change width of theinjection start timing of the pre-injection Q1 in the low-load firstsegment A1_a with respect to the engine speed. Note that in thisembodiment, as illustrated in FIG. 11, the injection start timing ti_13of the pre-injection Q1 when the engine speed is the third speed N3 isslightly retarded from the injection start timing ti_11 of thepre-injection Q1 when the engine speed is lower than the first speed N1.

As illustrated in FIG. 11, in the high-load first segment A1_b as awhole, the injection start timing of the middle injection Q2 is set toretard as the engine speed increases.

Specifically, within the engine speed range below the first speed N1,the injection start timing of the middle injection Q2 is set to a timingti_14 on the intake stroke, fixed regardless of the engine speed. Whenthe engine speed exceeds the first speed N1, the injection start timingof the middle injection Q2 is gradually retarded from the timing ti_14as the engine speed increases. Within the engine speed range above thesecond speed N2, the injection start timing of the middle injection Q2is set near a timing ti_15 for when the engine speed is the second speedN2, substantially fixed regardless the engine speed (a timing retardedfrom the timing ti_14 for when the engine speed is the first speed N1).The injection start timing of the middle injection Q2 is set to IBDC atan engine speed between the first speed N1 and the second speed N2, andwhen the engine speed is the second speed N2, the injection start timingof the middle injection Q2 is on the compression stroke. Note that asillustrated in FIG. 11, in an engine speed range from the second speedN2 to the speed slightly higher than N2, the injection start timing ofthe middle injection Q2 is slightly advanced as the engine speedincreases. Note that this advancing amount is extremely small, and inthe engine speed range above the second speed N2, the injection starttiming of the middle injection Q2 is substantially fixed regardless ofthe engine speed.

As described above, in the high-load first segment A1_b, the postinjection Q3 is stopped when the engine speed is high. In thisembodiment, when the engine speed exceeds the first speed N1, the postinjection Q3 is stopped. In the high-load first segment A1_b and theengine speed range below the first speed N1, the injection start timingof the post injection Q3 is a timing ti_16 near CTDC, fixed regardlessof the engine speed.

(Ratio of Injection Amount)

FIG. 12 is a chart illustrating a change in the ratio of the injectionamount of each of the injections Q1, Q2, and Q3 in relation to theengine speed, in the high-load first segment A1_b.

As illustrated in FIG. 12, in the high-load first segment A1_b as awhole, the ratio of the injection amount of the pre-injection Q1 is setlarger as the engine speed increases.

Specifically, in the engine speed range below the first speed N1, theratio of the injection amount of the pre-injection Q1 is fixedregardless of the engine speed and is set to a ratio Ri_11 close to100%. When the engine speed exceeds the first speed N1, the ratio of theinjection amount of the pre-injection Q1 is gradually increased from theratio Ri_11 as the engine speed increases. Within the engine speed rangeabove the second speed N2, the ratio of the injection amount of thepre-injection Q1 is a ratio Ri_12 for when the engine speed is thesecond speed N2, fixed regardless of the engine speed (a value largerand closer to 100% than the ratio Ri_11 for when the engine speed is thefirst speed N1). For example, as described above, the ratio Ri_11 of theinjection amount of the pre-injection Q1 at the operation point P1_bwhen the engine speed is below the first speed N1 is about 80%, and theratio Ri_12 of the injection amount of the pre-injection Q1 at theoperation point P3 when the engine speed is above the second speed N2 isabout 90%.

On the other hand, as illustrated in FIG. 12, the ratio of the injectionamount of the middle injection Q2 is constantly fixed at a value Ri_13regardless of the engine speed in the high-load first segment A1_b. Inthe high-load first segment A1_b, the ratio of the injection amount ofthe pre-injection Q1 is constantly kept larger than that of the middleinjection Q2 regardless of the engine speed.

Further, as illustrated in FIG. 12, in the high-load first segment A1_b,the ratio of the injection amount of the post injection Q3 is setsmaller as the engine speed increases. That is, as described above, inthe high-load first segment A1_b, the post injection Q3 is stopped whenthe engine speed exceeds the first speed N1, the ratio of the injectionamount of the post injection Q3 when the engine speed is below the firstspeed N1 is set larger than 0% (e.g., about 10%), and the ratio of theinjection amount of the post injection Q3 when the engine speed exceedsthe first speed N1 is 0%.

(5) Swirl Control

Next, a swirl control within the first operating range A1 will bedescribed in detail.

(5-1) Opening Setting of Swirl Valve

FIG. 13 is a map illustrating a specific example of a target value ofthe opening of the swirl valve 18 set within the first operating rangeA1 (hereinafter, may be referred to as “target swirl valve opening”).FIG. 14 is a chart illustrating a change in the target swirl valveopening when the engine speed is changed while the engine load is fixed(along a line V2 of FIG. 13). As illustrated in FIGS. 13 and 14, withinthe first operating range A1, the target external EGR ratio is set toincrease substantially between 20 and 40% as the engine speed or load isincreased.

For example, the target swirl valve opening is uniformly set to 20%within a base segment c1 which is a lowest-speed, lowest-load segmentwithin the first operating range A1, and is set to gradually increase asthe engine speed or load increases within a main segment c2 where theengine speed or load is higher than the base segment c1. In the mainsegment c2, the target swirl valve opening approaches 20% as the enginespeed and load decrease closer to the base segment c1, and is increasedfrom 20% to approximately 40% at most as the engine speed and loadincrease to move away from the base segment c1. For example, when theengine speed increases to cross the base segment c1 and the main segmentc2 in this order (along the line V2 of FIG. 13), as illustrated in FIG.14, the target swirl valve opening is kept at 20% while the engine speedis within the base segment c1, and after shifting to the main segmentc2, it is increased substantially at a fixed rate as the engine speedincreases. In other words, in this embodiment, the change rate of theopening of the swirl valve 18 with respect to the engine speed (and thusa change rate of the intensity of the swirl flow) is set to increase asthe engine speed increases. Here, a boundary speed S lies between thebasic segment c1 and the main segment c2.

During the operation in the first operating range A1, the ECU 100controls the opening of the swirl valve 18 according to the engine speedand load in accordance with the target swirl valve opening set asdescribed above.

(5-2) Effect of Swirl Flow

The opening control of the swirl valve 18 as described above isperformed in order to control the distribution of the mixture gas insidethe combustion chamber 6 (adjust the fuel concentration difference) byusing the swirl flow. That is, since the swirl flow swirling around thecylinder axis Z brings a relatively large amount of fuel to the centerportion of the combustion chamber 6, by suitably adjusting the intensityof the swirl flow and the fuel injection start timing, a desired fuelconcentration difference in the radial directions of the combustionchamber 6 can be generated. Hereinafter, the above-described effect ofthe swirl flow will be described in detail while showing the behavior offuel (spray) when the fuel is injected on the intake stroke in a statewhere a sufficiently strong swirl flow is formed.

FIG. 15 shows diagrams illustrating the behavior of fuel (spray)injected by the injector 15, in relation to the swirl flow. Theperspective view at the left end of FIG. 15 schematically shows a stateof the combustion chamber 6 at a given timing on the intake stroke atwhich the volume of the combustion chamber 6 is relatively large. InFIG. 15, the swirl valve 18 is substantially closed. In this case, airis introduced into the combustion chamber 6 mainly from the first intakeport 9A, and a strong swirl flow (lateral vortex) swirling around thecylinder axis Z is formed as indicated by an arrow in FIG. 15. Thisswirl flow is, as described before, an inclined swirl flow mixed withthe tumble flow (vertical vortex).

Here, in a direction perpendicular to the intake and exhaust directions,the side of the combustion chamber 6 at which the swirl valve 18 isprovided (the second intake port 9B side) is a front side, the sidewhere the swirl valve 18 is not provided (the first intake port 9A side)is a rear side. The swirl flow formed by closing the swirl valve 18(that is, the inclined swirl flow) flows upwardly at the exhaust side ofthe combustion chamber 6 from the first intake port 9A, then passes afront section of the combustion chamber 6 obliquely downwardly whileswirling in a large circle, and reaches a lower section at the intakeside of the combustion chamber 6. Further, the swirl flow passes througha rear section of the combustion chamber 6 obliquely upwardly whileswirling in a large circle, and returns to the upper section at theexhaust side of the combustion chamber 6.

A reference character D in the perspective view at the left end of FIG.15 indicates a virtual plane bisecting an internal space of thecombustion chamber 6 in the front and rear directions of the engine(directions perpendicular to the intake and exhaust directions), andeach of the schematic views (a) to (j) located on the right side of theperspective view indicates states of the front and rear sides of thecombustion chamber 6 divided by the virtual plane D. For example, thefive parts (a) to (e) in the upper part of FIG. 15 illustrate in achronological order an influence of an upstream portion of the swirlflow flowing at the front side of the combustion chamber 6 on each sprayof the fuel. Further, the five parts (f) to (j) in the lower part ofFIG. 15 illustrate in a chronological order an influence of a downstreamportion of the swirl flow flowing at the rear side of the combustionchamber 6 on each spray of the fuel.

The outlined arrows in Parts (a) to (j) of FIG. 15 show a main stream ofthe inclined swirl flow generated inside the combustion chamber 6 (acenter portion of the flow with strong stream, hereinafter may simply bereferred to as “the swirl flow”). Note that the main stream of the swirlflow has therearound a weak side stream flowing in the same direction asthe main stream. Although the flow of the fuel spray may be influencedby the side stream, since the current direction of the side stream isthe same as the main stream and also the main stream is more intense,even when the fuel spray is influenced by the side stream, the mainstream has a dominant influence finally. Therefore, a later-describedphenomenon in which the mixture gas distribution is formed by the swirlflow rarely changes due to the side stream.

First, the behavior of fuel injected to the front side of the combustionchamber 6 will be described. Part (a) of FIG. 15 illustrates the frontside of the combustion chamber 6 immediately after the fuel is injectedfrom the injector 15. Although this fuel injection simultaneously formsa spray of five fuel portions f1 to f5 at the front side of thecombustion chamber 6, none of the spray has reached the swirl flow atthis point. Note that in FIG. 15, the spray of the fuel portions f1 tof5 is illustrated as arrows along the center lines L1 to L5 of therespective fuel portions (see Part (b) of FIG. 15 and FIG. 2).Similarly, the spray of the fuel portions f6 to f10 described later isalso illustrated as arrows along the center lines L6 to L10 of therespective fuel portions (see Part (g) of FIG. 15 and FIG. 2).

As illustrated in the Part (b) of FIG. 15, among all the fuel portionsf1 to f5 at the front side, the fuel portion f1 injected from the nozzleport closest (having the shortest reach distance) to the swirl flowreaches the swirl flow first. Next, as illustrated in Part (c) of FIG.15, the fuel portion f2 injected from the nozzle port having the secondshortest reach distance reaches the swirl flow. The fuel portion f2reaches the swirl flow on the downstream side of the position at whichthe fuel portion f1 precedingly reaches the swirl flow. On the otherhand, the fuel portion f1 moves downstream together with the swirl flow.Therefore, when the fuel portion f2 reaches the swirl flow, it joinswith the fuel portion f1 moving together with the swirl flow.

As illustrated Part (d) of FIG. 15, the fuel portion f3 injected fromthe nozzle port having the third shortest reach distance reaches theswirl flow on the downstream side of the position at which the fuelportion f2 reaches the swirl flow. Here, the fuel portion f3 joins withthe fuel portions f1 and f2 that already joined to flow together withthe swirl flow.

Next, as illustrated in Part (e) of FIG. 15, the fuel portion f4injected from the nozzle port having the fourth shortest reach distancereaches the swirl flow. In this example, the fuel portion f4 reaches theswirl flow in a lower end section of the combustion chamber 6. Here, thefuel portion f4 joins with the fuel portions f1, f2, and f3 that alreadyjoined to flow together with the swirl flow.

Further, as illustrated in Part (d) of FIG. 15, the fuel portion f5adjacent to the fuel portion f4 (closest to the intake side) reaches awall surface 6 a of the combustion chamber 6 first. As illustrated inParts (d) and (e) of FIG. 15, the fuel portion f5 reached the wallsurface 6 a reaches the swirl flow by moving downward along the wallsurface 6 a. Here, the fuel portion f5 joins with the fuel portions f1,f2, f3, and f4 that already joined to flow together with the swirl flow.

As described above, in this embodiment, since the plurality of fuelportions (the spray of the fuel portions f1 to f5) are radially injectedfrom the injector 15 to the front side of the combustion chamber 6 wherethe swirl flow is formed, the fuel portion which reaches the swirl flowfirst (e.g., the fuel portion f1) moves downstream along the swirl flowand then joins with another fuel portion (e.g., the fuel portion f2)which arrives following the swirl flow. In this embodiment, all the fuel(the fuel portions f1 to f5) injected from the injector 15 to the frontside of the combustion chamber 6 join on the swirl flow by thismechanism. This leads to forming a rich mixture gas at a high fuelconcentration.

Next, the behavior of the fuel injected to the rear side of thecombustion chamber 6 will be described. Part (f) of FIG. 15 in the lowerpart illustrates a state of the rear side of the combustion chamber 6immediately after the fuel is injected from the injector 15. Althoughthis fuel injection simultaneously forms spray of five fuel portions f6to f10 at the rear side of the combustion chamber 6 (also simultaneouslyto the spray of the fuel portions f1 to f5 at the front side), none ofthe spray reaches the swirl flow at this point.

As illustrated in Part (g) of FIG. 15, among all the fuel portions f6 tof10 at the rear side, the fuel portion f10 injected from the nozzle portclosest to the swirl flow reaches the swirl flow first. Next, asillustrated in Part (h) of FIG. 15, the fuel portion f9 injected fromthe nozzle port having the second shortest reach distance reaches theswirl flow. The fuel portion reaches the swirl flow on the upstream sideof the position at which the fuel portion f10 precedingly reaches theswirl flow.

When the fuel portion f9 reaches the swirl flow, the fuel portion f10that previously reached the swirl flow has already moved downstreamtogether with the swirl flow to some extent. That is, the fuel portionf10 moves away from the fuel portion during the time between the arrivalof the fuel portion f10 at the swirl flow and the arrival of the fuelportion at the swirl flow thereafter, and does not join with the fuelportion f9. After the fuel portion f9 reaches the swirl flow, the fuelportion f9 moves downstream together with the swirl flow, while the fuelportion f10 moves even further downstream. Therefore, in either case thefuel portion f9 does not join with the fuel portion f10. In this manner,the fuel portions f9 and f10 move along the swirl flow while stayingseparated from each other.

Next, as illustrated in Part (i) of FIG. 15, the fuel portion f8injected from the nozzle port at the third shortest reach distancereaches the swirl flow on the upstream side of the position at which thefuel portion f9 reaches the swirl flow. Here, since the fuel portions f9and f10 which precedingly reach the swirl flow move downstream togetherwith the swirl flow, these fuel portions f9 and f10 do not join with thefuel portion f8.

As described above, in this embodiment, since the plurality of fuelportions (the spray of the fuel portions f8 to f10) are radiallyinjected from the injector 15 to the rear side of the combustion chamber6 where the swirl flow is formed, the fuel portion which reaches theswirl flow first (e.g., the fuel portion f10) does not join with thefuel portions that reach the swirl flow later (e.g., the fuel portionsf8 and f9) which arrive following the swirl flow. In this embodiment,approximately 30% of the fuel injected by the injector 15 diffuses,which leads to forming a homogeneous mixture gas with the fuel spreadingwidely and thin.

In the fuel injected to the rear side of the combustion chamber 6 (thefuel portions f6 to f10), the fuel other than the fuel portions f8 tof10 separated from each other, that is, the fuel portions f6 and f7,join with the fuel portions f1 to f5 injected to the front side of thecombustion chamber 6.

For example, among the fuel portions f6 to f10 at the rear side, thefuel portion f7 injected from the nozzle port at the fourth shortestreach distance reaches the swirl flow in the lower end portion of thecombustion chamber 6 as illustrated in Part (j) of FIG. 15. In the lowerend portion of the combustion chamber 6, since the fuel portions f1 tof5 at the front side join together by the mechanism described above (seePart (e) of FIG. 15), the fuel portion f7 joins with the fuel portionsf1 to f5 which join precedingly.

As illustrated in Part (i) of FIG. 15, the fuel portion f6 adjacent tothe fuel portion f7 (closest to the intake side) reaches the wallsurface 6 a of the combustion chamber 6 first. As illustrated in Parts(i) and (j) of FIG. 15, the fuel portion f6 reached the wall surface 6 areaches the swirl flow by moving downward along the wall surface 6 a.Here, the fuel portion f6 joins with the fuel portion f7 and the fuelportions f1 to f5 at the front side. Thus, in this embodiment,approximately 70% of the fuel injected by the injector 15 joinstogether.

The perspective view at the right end of FIG. 15 schematicallyillustrates a state of the combustion chamber 6 immediately after allthe fuel (the fuel portions f1 to f10) injected from the injector 15reach the swirl flow. As illustrated in this figure, in this embodimentwhere a major portion (70%) of the fuel injected from the injector 15joins together on the swirl flow, the (rich) mixture gas at asufficiently high fuel concentration is formed along the swirl flow. Therich mixture gas gradually moves toward the center of the combustionchamber 6 while moving circumferentially within the combustion chamber 6together with the swirl flow.

FIG. 16 shows diagrams illustrating the mixture gas moving with theswirl flow within the combustion chamber 6 seen from above. Asillustrated in the figure, the swirl flow formed inside the combustionchamber 6 grows sufficiently as the intake stroke progresses, thenattenuates by receiving air resistance, and moves toward the center ofthe combustion chamber 6 while gradually diffusing.

Part (a) of FIG. 16 illustrates a state where the spray of the fuelportions f1 to f7 joins together to form a rich mixture gas (areas withfine dots). As indicated by the arrows in this drawing, this richmixture gas moves together with the swirl flow which changes asdescribed above, and thus moves toward the center of the combustionchamber 6 while gradually diffusing. As a result, as illustrated in Part(b) of FIG. 16, the relatively rich mixture gas is unevenly distributedin the center portion of the combustion chamber 6 immediately before thecombustion starts.

Part (c) of FIG. 16 illustrates a state where (lean) mixture gas at alow fuel concentration is formed by the fuel portions f8 to f10 which donot join with each other. As indicated by the arrows in the drawing,this lean mixture gas moves together with the swirl flow to move towardthe center of the combustion chamber 6 while sufficiently diffusing.Thus, as illustrated in Part (d) of FIG. 16, immediately before thecombustion starts, relatively lean mixture gas spreading in the entirecombustion chamber 6 is formed.

Part (e) of FIG. 16 illustrates a state where the mixture gasillustrated in Parts (b) and (d) are superimposed. As illustrated inthis figure, the mixture gas formed by the fuel portions f1 to f7 joinedtogether (Part (b) of FIG. 16), is mixed with the mixture gas formed bythe diffusing fuel portions f8 to f10 (Part (d) of FIG. 16), and thus astratified mixture gas in which the fuel concentration is higher in thecenter portion than in the outer circumferential portion is formed inthe combustion chamber 6. That is, the rich mixture gas with relativelyhigh fuel concentration is formed in the center portion of thecombustion chamber 6 and the lean mixture gas with relatively low fuelconcentration is formed in the outer circumferential portion of thecombustion chamber 6.

With the effect of the swirl flow as described above, even when the fuelis injected at a relatively early timing of the intake stroke, themixture gas at the timing of spark ignition (at the start of the SPCCIcombustion) is stratified to some extent, which means that relativelyrich mixture gas is formed in the center portion of the combustionchamber 6. Such stratification of the mixture gas becomes more obviousas the swirl flow is enhanced.

For example, at the operation point P1_a of the first operating rangeA1, the major portion of the fuel is injected on the intake stroke asdescribed above (see Part (a) of FIG. 8). However, at a point at whichthe engine speed and load are sufficiently low as at the operation pointP1, since the swirl valve 18 is set to have the lowest opening (20%) anda sufficiently strong swirl flow is formed as illustrated in FIG. 8,even when the major portion of the fuel is injected on the intake strokeas described above, the fuel distribution is not made even and the fuelconcentration in the center portion of the combustion chamber 6 becomeshigher than in the outer circumferential portion. Moreover, at theoperation point P1_a, since additional fuel is injected in the finalstage of the compression stroke, this injected fuel is added to thecenter portion of the combustion chamber 6, resulting in more obviousstratification of the mixture gas. As a result, at the operation pointP1_a, sufficiently stratified mixture gas is formed in which A/F isbetween 20:1 and 30:1 in the center portion and 35:1 or larger in theouter circumferential portion. Such stratification of the mixture gasacts advantageously on the growth of the flame after the spark ignition.That is, while the spark ignition by the spark plug 16 acts on themixture gas in the center portion of the combustion chamber 6 to form aflame core, since the A/F in the center portion is relatively rich asdescribed above, the growth of the flame core is stimulated and theprogression of subsequent combustion is stabilized.

(6) About SI Ratio

As described above, in this embodiment the SPCCI combustion combined theSI combustion and the CI combustion is performed within the first andsecond operating ranges A1 and A2. In this SPCCI combustion, it isimportant to control the ratio of the SI combustion with the CIcombustion according to the operating condition.

In this embodiment, as this ratio, an SI ratio which is a ratio of aheat generation amount by the SI combustion to the total heat generationamount by the SPCCI combustion (the SI combustion and the CI combustion)is used. The flection point X in the waveform of FIG. 6 is a flectionpoint appearing when the combustion mode switches from the SI combustionto the CI combustion and the crank angle θci corresponding to thisflection point X may be defined as the start timing of CI combustion.Further, an area R1 of the waveform of the heat generation rate locatedon the advance side of this θci (the start timing of CI combustion) isset as the heat generation amount by the SI combustion, and an area R2of the waveform of the heat generation rate located on the retardingside of θci is the heat generation rate by the CI combustion. Thus, theSI ratio defined by (heat generation amount by SI combustion)/(heatgeneration amount by SPCCI combustion) may be expressed by R1/(R1+R2)using the respective areas R1 and R2. That is, in this embodiment, theSI ratio=R1/(R1+R2).

In the case of CI combustion, since the mixture gas combusts a pluralityof times simultaneously by self-ignition, the heat generation rateeasily increases compared to the SI combustion which is caused by flamepropagation, and loud noise easily occurs. Therefore, as a whole, the SIratio (=R1/(R1+R2)) in the SPCCI combustion is desirably increased asthe engine load increases. This is because, when the load is high, thefuel injection amount is larger and the total heat generation amount inthe combustion chamber 6 is larger compared to when the load is low, andtherefore, loud noise is generated if the SI ratio is reduced (that is,the proportion of CI combustion is increased). Conversely, the CIcombustion is excellent in terms of thermal efficiency. Therefore, it ispreferable to apply the CI combustion to the largest fuel amountpossible unless the noise becomes an issue. Therefore, as a whole, theSI ratio in the SPCCI combustion is desirably reduced as the engine loaddecreases (i.e., the proportion of the CI combustion is increased). Inview of these points, in this embodiment, the SI ratio aimed for (targetSI ratio) is determined in advance so that the SI ratio increases as theengine load becomes higher (in other words, the SI ratio decreases asthe engine load becomes lower). Accordingly, in this embodiment, atarget θci which is a start timing of the CI combustion when thecombustion conforming to this target SI ratio is performed is alsodetermined in advance for each operating condition of the engine.

In order to achieve the target SI ratio and the target θci, the controlamounts, such as the ignition timing, the fuel injection amount andinjection start timing, and the EGR ratio (the external EGR ratio andthe internal EGR ratio), need to be adjusted for each operatingcondition. For example, as the ignition timing is advanced, a largeramount of fuel is combusted in the SI combustion, and the SI ratioincreases. Further, as the injection start timing of the fuel isadvanced, a larger amount of fuel is combusted in the CI combustion, andthe SI ratio decreases. Alternatively, as the temperature of thecombustion chamber 6 rises according to the increase of the EGR ratio, alarger amount of fuel is combusted in the CI combustion, and the SIratio decreases. Moreover, since a change in the SI ratio is followed bya change in the θci, changes in these control amounts (the ignitiontiming, the injection start timing, the EGR ratio, etc.) are elementsfor adjusting the θci.

Based on such a tendency, in this embodiment, the target values of theignition timing, the fuel injection amount and injection start timing,the external EGR ratio, the valve operation timings (and thus theinternal EGR ratio), etc. are determined in advance for each operatingcondition so as to achieve a value combination achievable of the targetSI ratio and the target θci described above. In the operation with theSPCCI combustion (i.e., in the operation within the first and secondoperating ranges A1 and A2), the ECU 100 controls the injector 15, thespark plug 16, the EGR valve 53, the intake and exhaust VVTs 13 a and 14a, etc. based on the target values of these control amounts. Forexample, the spark plug 16 is controlled based on the target value ofthe ignition timing and the injector 15 is controlled based on thetarget values of the fuel injection amount and injection start timing.Further, the EGR valve 53 and the intake and exhaust VVTs 13 a and 14 aare controlled based on the respective target values of the external EGRratio in the combustion chamber 6 and the valve operation timings(internal EGR ratio), and the recirculation amount of exhaust gas(external EGR gas) through the EGR passage 51 and the residual amount ofburned gas (internal EGR gas) by the internal EGR are adjusted.

(7) Control based on Target θci

As described above, in this embodiment, the target values of the controlamounts, such as the ignition timing, the fuel injection amount andinjection start timing, and the EGR ratio, are determined in advance sothat the SI ratio and θci match their respective target values (thetarget SI ratio and the target θci); however, controlling the controlamounts (the ignition timing, etc.) to the target values do notguarantee to achieve the target SI ratio and the target θci. Therefore,in this embodiment, the start timing of the CI combustion, θci, isidentified for each combustion cycle and the ignition timing is adjustedbased on the identified θci.

FIG. 17 is a flowchart illustrating details of the control executed bythe ECU 100 in the SPCCI combustion, that is, in the operation withinthe first and second operating ranges A1 and A2 illustrated in FIG. 5.When the control illustrated in this flowchart starts, at S1, the ECU100 determines the fuel injection amount from the injector 15 and itsinjection start timing based on the engine speed detected by the crankangle sensor SN1 and the engine load identified by the detection valueof the accelerator sensor SN9 (accelerator opening), the detection valueof the airflow sensor SN3 (intake flow rate), etc. Note that theinjection amount and injection start timing of the fuel determined hereare the injection amount and injection start timing defined in advancefor each operating condition of the engine so that the target θci isachieved.

Next, the ECU 100 shifts to S2 to determine whether the intake valve 11is closed (switched from “open” to “closed”). In this embodiment, sincethe valve operation timing is determined in advance for each engineoperating condition, IVC is also known in advance for each operatingcondition. The ECU 100 determines whether IVC determined for eachoperating condition has passed based on the detection value of the crankangle sensor SN1, and determines that the intake valve 11 is closed whendetermined that IVC has passed.

If S2 is YES and it is confirmed that the intake valve 11 is closed, theECU 100 shifts to S3 to estimate the EGR ratio and the in-cylindertemperature (the internal temperature of the combustion chamber 6). Forexample, the ECU 100 estimates the EGR ratio (the external EGR ratio andthe internal EGR ratio) based on various parameters including thein-cylinder pressure detected by the in-cylinder pressure sensor SN2when the intake valve 11 is closed (that is, IVC), the flow rate of theintake air (fresh air) detected by the air flow sensor SN3 before IVC,and the pressure difference between upstream and downstream of the EGRvalve 53 detected by the pressure difference sensor SN8 before IVC.Further, the in-cylinder temperature is estimated based on the estimatedEGR ratio and the detection value detected by the second intaketemperature sensor SN6.

Note that as described above, in this embodiment, the respective targetvalues of the external EGR ratio and the valve operation timing(internal EGR ratio) are determined in advance, and the opening of theEGR valve 53 and the intake and exhaust VVTs 13 a and 14 a areaccordingly controlled. However, the actual external EGR ratio and theinternal EGR ratio can be thought to vary in different circumstances,and such EGR ratio fluctuation or the accompanying variation in thein-cylinder temperature influence the start timing of the CI combustion(θci) and the SI ratio. Therefore, in order to adjust the ignitiontiming considering such influence, in this embodiment, at S3 the actualEGR ratio and the in-cylinder temperature are estimated for eachcombustion cycle.

Next, the ECU 100 shifts to S4 to determine whether the current crankangle is a specific crank angle based on the detection value of thecrank angle sensor SN1. This specific crank angle is determined inadvance as the timing for determining the ignition timing by the sparkplug 16, and determined to be a suitable timing from IVC to CTDC (e.g.,about 60° CA before CTDC).

If S4 is YES and it is confirmed that the current crank angle is thespecific crank angle, the ECU 100 shifts to S5 to determine the ignitiontiming for achieving the target θci.

For example, at S5, the ECU 100 determines the ignition timing forachieving the target θci based on an initial target value of theignition timing determined corresponding to the target θci (hereinafter,referred to as “default ignition timing”) and the estimated values ofthe EGR ratio and the in-cylinder temperature obtained at S3.

In this embodiment, in order to determine the ignition timing, a modelformula is prepared in advance for determining a correction amount ofthe ignition timing based on the estimated EGR ratio and in-cylindertemperature. For example, the ignition timing needs to be deviatedgreater from the default ignition timing as the estimated EGR ratio andin-cylinder temperature deviate greater from their target values. On theother hand, in this embodiment, since the initial target values areadopted as they are as the fuel injection amount and injection starttiming as described at S1, the deviations in the fuel injection amountand injection start timing from the target values are not required to betaken into consideration. Therefore, in this embodiment, an arithmeticformula having the deviation amounts of the EGR ratio and in-cylindertemperature from the target values as an input element and thecorrection amount of the ignition timing as an output element isprepared in advance as the model formula. In other words, by inputtingthe deviation amounts of the EGR ratio and in-cylinder temperature fromthe target values into the model formula, the correction amount of theignition timing required to match θci with the target θci (thecorrection amount from the default ignition timing) is obtained. At S5,a timing retarded or advanced from the default ignition timing by thecorrection amount calculated based on the model formula is determined asthe final ignition timing. Note that when the EGR ratio and in-cylindertemperature estimated at S3 are the same as the target values, thedefault ignition timing is adopted as the ignition timing as it is.

Next, the ECU 100 shifts to S6 to cause the spark plug 16 to ignite atthe ignition timing determined at S5, so as to trigger the SPCCIcombustion of the mixture gas.

Next, the ECU 100 shifts to S7 to calculate θci (the start timing of theCI combustion) based on the in-cylinder pressure detected over a givenperiod overlapping with the combustion period of the SPCCI combustion(the period from the start until the end of combustion), and correct themodel formula used at S5 based on the calculated θci. That is, the ECU100 calculates a heat generation rate accompanying the combustion foreach crank angle based on the waveform of the in-cylinder pressuredetected by the in-cylinder pressure sensor SN2 during the given period,and the timing when the SI combustion is switched to the CI combustion(the timing corresponding to the flection point X of FIG. 6) isidentified as θci based on the data of the heat generation rate for eachcrank angle. Further, a deviation amount between this identified θci andthe target θci is calculated and, based on this calculated deviationamount, the model formula for determining the ignition timing (theignition timing for achieving the target θci) is corrected. Thiscorrection of the model formula leads to an improvement of accuracy indetermining the ignition timing next time in a similar condition.

(8) Operations and Effects

As described above, in this embodiment, in the low-load first segmentA1_a of the first operating range A1 where the SPCCI combustion isperformed in the A/F lean environment, the ratio of the injection amountof the post injection Q3 is increased as the engine speed increases.Therefore, there is an advantage in that high combustibility is ensuredregardless of the engine speed.

That is, since the moving speed of the piston 5 is higher as the enginespeed increases, the combustion chamber 6 sharply expands after thespark ignition, and as a result, the growth of the flame core is easilyinterrupted. On the other hand, in this embodiment, at the high speedside in the low-load first segment A1_a, the ratio of the injectionamount of the post injection Q3 is increased than the lower speed side.Therefore, when the engine speed is high, at the ignition timing nearCTDC, the fuel concentration of the mixture gas locally formed insidethe combustion chamber 6 is increased. Thus, even under a conditionwhere the engine speed is high and a rate of an expansion inside thecombustion chamber after the spark ignition is high (as a result, theflame core is difficult to grow), the flame core is formed and itsgrowth is stimulated so that the SI combustion progresses stably and thesubsequent CI combustion is surely started, which avoids the starttiming of the CI combustion (θci) from greatly varying between cycles.As described above, according to this embodiment, stable partialcompression-ignition combustion is achieved regardless of the enginespeed.

Further, in this embodiment, at the high-speed side in the low-loadfirst segment A1_a, the pre-injection Q1 and the post-injection Q3 areperformed on the compression stroke.

Therefore, compared to a case where only the pre-injection Q1, or boththe pre-injection Q1 and the post-injection Q3, are performed on theintake stroke, the mixture gas having the high fuel concentration issurely formed in a center portion of the combustion chamber 6 by thepre-injection Q1 and the post-injection Q3. In this manner, the flamecore is formed and grown more reliably when the engine speed is high.

For example, in this embodiment, the fuel injected on the intake strokeby the swirl flow is gathered in the center portion of the combustionchamber. However, as described above, the swirl flow grows as the intakestroke progresses, then attenuates due to air resistance, and graduallydiffuses. Therefore, when the engine speed is high and the mixture gashaving the high fuel concentration needs to be formed in the centerportion of the combustion chamber 6 (the fuel concentration of themixture gas in the center portion of the combustion chamber 6 needs tobe increased), the effect of the swirl flow alone may be insufficient.In this regard, in this embodiment, at the high speed side in thelow-load first segment A1_a, with the combination of the effect of theswirl flow with the effect of performing the pre-injection Q1 and thepost-injection Q3 on the compression stroke to prevent the fuel spray ofthese injections from diffusing to the entire combustion chamber 6 byCTDC, the mixture gas having the high fuel concentration is formed inthe center portion of the combustion chamber 6 more reliably (the fuelconcentration of the mixture gas in the center portion of the combustionchamber 6 is increased more).

Here, when the engine speed is low, the expansion of the combustionchamber 6 after the spark ignition is relatively slower, and the flamecore is easily formed and grown. Thus, even if the fuel amount suppliedto the combustion chamber 6 on the compression stroke is reduced byperforming the pre-injection Q1 on the intake stroke, the SPCCIcombustion is suitably achieved when the engine speed is low.

In this regard, in this embodiment, when the engine speed is low in thelow-load first segment A1_a, the injector 15 is controlled to performthe pre-injection Q1 on the intake stroke and the post-injection Q3 onthe compression stroke.

Further, in this embodiment, during the operation within the firstoperating range A1, in the state where the opening of the swirl valve 18is reduced to a relatively low opening (about 20 to 40%), that is, in astate where a relatively strong swirl flow is formed inside thecombustion chamber 6, the fuel is radially injected at a suitable timing(the timing when the spray of the fuel reaches the swirl flow) from themulti-hole injector 15 located at the center portion of the ceilingsurface of the combustion chamber 6, and thus the stratified mixture gasin which the fuel concentration is higher in the center portion of thecombustion chamber 6 than in the outer circumferential portion is formed(see FIGS. 15 and 11). With this configuration, a flame core is reliablyformed by the spark ignition using the spark plug 16 within relativelyrich mixture gas formed in the center portion of the combustion chamber6, and the SI combustion and the subsequent CI combustion arestabilized.

Especially in this embodiment, the inclined swirl flow which flowsnonparallel to a plane perpendicular to the cylinder axis Z is formed asthe swirl flow, and the position and orientation of each nozzle port ofthe injector 15 are set so that the fuel portion which reaches the swirlflow first (e.g., the fuel portion f1) moves downstream along the swirlflow and then joins with another fuel portion which reaches the swirlflow later (e.g., the fuel portion f2). Thus, the (relatively rich)mixture gas at the high fuel concentration is reliably formed in thecenter portion of the combustion chamber 6 which is the finaldestination of the swirl flow, and therefore, the stratification of themixture gas described above is further stimulated.

Further in this embodiment, in the low-load first segment A1_a, theinjector 15 is controlled so that the ratio of the injection amount ofthe pre-injection Q1 is reduced when the engine speed increases and theratio of the injection amount of the post-injection Q3 is increased whenthe engine speed increases.

Thus in the low-load first segment A1_a, the fuel concentration of themixture gas locally formed in the combustion chamber 6 at the ignitiontiming is increased furthermore reliably, and the CI combustion is moresurely performed.

Further in this embodiment, in the high-load first segment A1_b, theinjector 15 is controlled so that the swirl flow is generated and theratio of the injection amount of the post-injection Q3 is reduced whenthe engine speed increases.

According to this configuration, in the high-load first segment A1_b,the suitable CI combustion is achieved more surely. That is, when theengine load is high, the fuel amount injected into the combustionchamber increases, and when the engine speed is high, a time lengthcorresponding to one crank angle becomes short. Therefore, if the ratioof the fuel injection amount of the post-injection Q3 is increased whenthe engine load and speed are high, the fuel is not sufficientlyvaporized until the ignition timing and soot, etc. may increase.Therefore, according to this configuration, at the higher speed side inthe high-load first segment A1_b where the engine load and speed arerelatively high, the suitable CI combustion is achieved while reducingthe generation of soot, etc.

Further in this embodiment, within the first operating range A1, thecontrol for performing the SPCCI combustion while forming the A/F leanenvironment in which A/F, which is the mass ratio of air to the fuelinside the combustion chamber 6, is larger than the stoichiometricair-fuel ratio, is executed. Therefore, within the first operating rangeA1, the SPCCI combustion is performed in an environment in which a heatcapacity ratio of the mixture gas is large and which is advantageous forthe thermal efficiency, and fuel efficiency of the engine is effectivelyimproved.

That is, when the engine is assumed to be an Otto cycle engine, itstheoretical thermal efficiency is determined based only on thecompression ratio of the engine and the heat capacity ratio of themixture gas, and it becomes large when either of the compression ratioand the heat capacity ratio is large. On the other hand, the heatcapacity ratio is known to decrease as the number of atoms constitutinga gas molecule within the mixture gas increases. For example, triatomicmolecules such as CO₂ and H₂O, which are contained within EGR gas bylarge amounts have a smaller heat capacity ratio than diatomic moleculessuch as N₂ and O₂, which are mainly contained within air (fresh air).From this, when the A/F lean environment is formed within the firstoperating range A1 and a larger amount of air than the air amountequivalent to the stoichiometric air-fuel ratio is introduced into thecombustion chamber 6, the ratio of the diatomic molecule within themixture gas may be increased to raise the heat capacity ratio, and thusthe thermal efficiency improves.

Further in this embodiment, when the SPCCI combustion is performed(during the operation within the first and second operating ranges A1and A2), the ignition timing by the spark plug 16 is adjusted so thatθci which is the start timing of the CI combustion matches with a giventarget θci for each operating condition of the engine, in other words,so that the SI ratio correlated with θci (the ratio of the heatgeneration amount by the SI combustion with respect to the total heatgeneration amount) matches with the target SI ratio. Therefore, theratio of the CI combustion is increased (i.e., the SI ratio is reduced)as much as possible within the range that the combustion noise does notbecome excessive, and the thermal efficiency by the SPCCI combustion isimproved as much as possible.

(9) Modifications

In this embodiment, the case where the three injections are performed inthe low-load first segment A1_a is described; however, the number ofinjections in the low-load first segment A1_a is not limited to this.

Moreover, the number of times injections are performed at the low speedside and the number of times injections are performed at the high speedside of the high-load first segment A1_b are not limited to thosedescribed above (three and two times). For example, also at thehigh-speed side of the high-load first segment A1_b, injections may beperformed the same number of times as at the low speed side (e.g., threetimes). However, as described above, by reducing the number of timesinjections are performed at the high speed side of the high-load firstsegment A1_b to be smaller than that at the low speed side, whileforming suitable mixture gas at the low speed side, the formation of themixture gas having excessively high fuel concentration at the high speedside is prevented.

Further, for each of the low speed side and high speed side of thelow-load first segment A1_a and the low speed side and high speed sideof the high-load first segment A1_b, the specific injection timings andthe ratios of the injection amount of the injections Q1, Q2, and Q3 arenot limited to those described above.

Although in this embodiment, the booster 33 mechanically driven by theengine body 1 is provided in the intake passage 30, instead of such amechanical booster 33 (supercharger), an electric booster driven by amotor, or a turbocharger driven by the energy of the exhaust gas may beprovided.

Although in this embodiment, the cavity 20 having a donut shape in theplan view to surround the conical-shaped bulge portion 20 a is formed inthe crown surface of the piston 5, the concave portion of the cavity 20facing the spark plug 16, that is, the concave portion located on theintake side of the bulge portion 20 a, may be formed so as to be smallerthan the concave portion on the opposite side (exhaust side). In thismanner, when the fuel is injected from the injector 15 in the finalstage of the compression stroke, fuel spray is moved to the vicinity ofthe electrode of the spark plug 16 more promptly.

Although in this embodiment, the intake and exhaust VVTs 13 a and 14 aare controlled to form the valve overlap period in which both the intakeand exhaust valves 11 and 12 are opened over TDC of the exhaust strokewhen performing the internal EGR in which the burned gas is left in thecombustion chamber 6, conversely, the internal EGR may be performed byforming a so-called negative overlap period in which both the intake andexhaust valves 11 and 12 are closed over TDC of the exhaust stroke.

In this embodiment, the intake VVT 13 a (exhaust VVT 14 a) adopts thevariable phase mechanism which simultaneously changes the open and closetimings of the intake valve 11 (exhaust valve 12); however, the intakeVVT 13 a may be a variable mechanism of a type which only changes theopen timing while fixing the close timing of the intake valve 11, andthe exhaust VVT 14 a may be a variable mechanism of a type which onlychanges the close timing while fixing the open timing of the exhaustvalve 12.

In this embodiment, the swirl valve 18 is provided to one of the twointake ports 9A and 9B (the second intake port 9B) provided to onecylinder 2, and the opening of the swirl valve 18 is changed to adjustthe strength of the swirl flow. However, the method of adjusting thestrength of the swirl flow with a swirl mechanism is not limited tothis. For example, the strength of the swirl flow may be adjusted bysetting the lift amount of the intake valve 11 which opens and closesthe first intake port 9A and the lift amount of the intake valve 11which opens and closes the second intake port 9B different from eachother or the operating timings of these two intake valves 11 differentfrom each other. In this manner, the intake valves 11 may be consideredthe “swirl mechanism.”

In this embodiment, for each operating condition (speed/load) of theengine, the optimal SI ratio (target SI ratio) and the correspondingstart timing of the CI combustion (target θci) with which the combustionnoise and the thermal efficiency have a well-balanced relationship aredetermined in advance, and during the SPCCI combustion, the ignitiontiming for achieving the target SI ratio and the target θci (thecorrection amount from the target value) is obtained using the modelformula based on the in-cylinder state functions at the given timingbefore the ignition (EGR ratio, temperature, etc.); however,alternatively/additionally to the ignition timing, the fuel injectiontiming from the injector 15 may be adjusted. Alternatively, both thefuel injection timing and amount may be adjusted.

In this embodiment, during the SPCCI combustion, the start timing of theCI combustion (θci) is calculated based on the waveform of thein-cylinder pressure detected by the in-cylinder pressure sensor SN2,and based on the calculated θci, the model formula is modified tocorrect the ignition timing; however, alternatively/additionally to θci,the SI ratio which is the ratio of the SI combustion may be calculatedand the model formula may be corrected based on the SI ratio.

Note that when calculating the SI ratio of each combustion as describedabove, various specific methods can be considered to calculate this SIratio.

For example, the heat generation rate at each crank angle timing may becalculated from the detected waveform by the in-cylinder pressure sensorSN2, and the areas R1 and R2 illustrated in FIG. 6 may be calculatedbased on the data (waveform) of the calculated heat generation rate. Inthis case, although the SI ratio may be calculated as SIratio=R1/(R1+R2) as described above, alternatively, SI ratio=R1/R2 maybe established.

Further, the SI ratio may be calculated using Δθ1 and Δθ2 illustrated inFIG. 18. That is, when the crank angle period of the SI combustion (thecrank angle period on the advancing side of the flection point X) is Δθ1and the crank angle period of the CI combustion (the crank angle periodon the retarding side of the flection point X) is Δθ2, SIratio=Δθ1/(Δθ1+Δθ2) or SI ratio=Δθ1/Δθ2 may be established.

Furthermore, when a peak of the heat generation rate of the SIcombustion is ΔH1 and a peak of the heat generation rate of the CIcombustion is ΔH2, SI ratio=ΔH1/(ΔH1+ΔH2) or SI ratio=ΔH1/ΔH2 may beestablished.

It should be understood that the embodiments herein are illustrative andnot restrictive, since the scope of the invention is defined by theappended claims rather than by the description preceding them, and allchanges that fall within metes and bounds of the claims, or equivalenceof such metes and bounds thereof, are therefore intended to be embracedby the claims.

DESCRIPTION OF REFERENCE CHARACTERS

1 Engine Body

6 Combustion Chamber

15 Injector

16 Spark Plug

18 Swirl Valve

40 Exhaust Passage

100 ECU (Controller)

What is claimed is:
 1. A control system for a compression-ignitionengine, comprising: an engine having a combustion chamber formed by acylinder, a piston, and a cylinder head; an injector attached to theengine and configured to supply fuel into the combustion chamber; aspark plug disposed to be oriented into the combustion chamber; a swirlvalve provided to an intake passage of the engine; and a controllerconnected to the injector, the spark plug, and the swirl valve, andconfigured to output a control signal to the injector, the spark plug,and the swirl valve, respectively, the controller including a processorconfigured to execute: a swirl adjusting module to adjust an opening ofthe swirl valve to generate a swirl flow inside the combustion chamber;a fuel injection amount controlling module to control fuel injectionamounts by the injector, of a pre-injection in which fuel is injectedinto the combustion chamber and a post-injection in which fuel isinjected again after the pre-injection, the fuel injection amountcontrolling module outputting a control instruction to the injector toincrease a ratio of an injection amount of the post-injection to a totalamount of fuel to be injected into the combustion chamber in one cycleas an engine speed increases; and a combustion controlling module tooutput an ignition instruction to the spark plug so as to ignite at agiven ignition timing after the swirl generation by the swirl adjustingmodule and the fuel injection by the fuel injection amount controllingmodule, so that partial compression-ignition combustion in which themixture gas combusts by flame propagation and then combusts bycompression ignition is performed.
 2. The control system of claim 1,wherein during the partial compression-ignition combustion in which theswirl flow is generated and the ratio of the injection amount of thepost-injection is increased as the engine speed increases, when theengine speed is higher than a given value, the fuel injection amountcontrolling module controls the injector to perform the pre-injectionand the post-injection on compression stroke.
 3. The control system ofclaim 1, wherein during the partial compression-ignition combustion inwhich the swirl flow is generated and the ratio of the injection amountof the post-injection is increased as the engine speed increases, thefuel injection amount controlling module controls the injector toperform the pre-injection on intake stroke and the pre-injection oncompression stroke when the engine speed is lower than a give value. 4.The control system of claim 1, wherein the injector at least has a firstnozzle port and a second nozzle port disposed at a center portion of aceiling surface of the combustion chamber and separated from each otherin a circumferential direction of the injector, the first and secondnozzle ports simultaneously injecting the fuel, wherein the swirl flowis an inclined swirl flow flowing nonparallel to a plane perpendicularto a center axis of the combustion chamber, and wherein the first andsecond nozzle ports are positioned and oriented so that a first fuelportion injected by the first nozzle port that has reached the swirlflow moves downstream along the swirl flow and then joins with a secondfuel portion injected by the second nozzle port that has reached theswirl flow.
 5. The control system of claim 1, wherein during the partialcompression-ignition combustion in which the swirl flow is generated andthe ratio of the injection amount of the post-injection is increased asthe engine speed increases, the fuel injection amount controlling modulecontrols the injector so that the ratio of an injection amount of thepre-injection to the total amount of fuel to be injected into thecombustion chamber in one cycle is reduced as the engine speedincreases, and the ratio of the injection amount of the post-injectionto the total amount of fuel to be injected into the combustion chamberin one cycle increases as the engine speed increases.
 6. The controlsystem of claim 1, wherein during a partial compression-ignitioncombustion in an operating range with a higher load than that of anoperating range where the partial compression-ignition combustion inwhich the swirl flow is generated and the ratio of the injection amountof the post-injection is increased as the engine speed increases, theswirl adjusting module causes the swirl valve to generate the swirl flowand the fuel injection amount controlling module controls the injectorso that the pre-injection and the post-injection are performed and theratio of the injection amount of the post-injection to the total amountof fuel to be injected into the combustion chamber in one cycle isreduced as the engine speed increases.
 7. A control system for acompression-ignition engine including an injector configured to supplyfuel into a combustion chamber and a spark plug configured to ignite amixture gas containing fuel supplied from the injector and air, theengine performing partial compression-ignition combustion includingspark ignition (SI) combustion performed by combusting a portion of themixture gas through spark ignition by the spark plug, followed bycompression ignition (CI) combustion performed by causing the remainingmixture gas to self-ignite, comprising: a swirl mechanism configured togenerate a swirl flow; and a processor configured to, during the partialcompression-ignition combustion: control the swirl mechanism to generatethe swirl flow; control the injector to perform a pre-injection in whicha given amount of fuel is injected into the combustion chamber and apost-injection in which fuel is injected again after the pre-injection;and increase a ratio of the injection amount of the post-injection to atotal amount of fuel to be injected into the combustion chamber in onecycle as the engine speed increases.